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Steam Power Plant 
Engineering 



BY 

G. F. GEBHARDT 

PROFESSOR OF MECHANICAL ENGINEERING, ARMOUR 

INSTITUTE OF TECHNOLOGY 

CHICAGO, ILL. 



FOURTH EDITION, REWRITTEN AND RESET 
TOTAL ISSUE SEVEN THOUSAND 



NEW YORK 

JOHN WILEY & SONS, Inc. 

London: CHAPMAN & HALL, Limited 

1913 






Copyright, 1908, 1910, 1913, 

BY 
G. F. GEBHARDT 






V 



V 



Stanbopc Press 

F. H. GILSON COMPANY 
BOSTON, U.S. A 



/«« 



©CU350258 



PREFACE TO FOURTH EDITION 



This work was first published in 1908. Many additions and changes 
were made in 1'909 and in 1911. The development of Steam Power 
Plants has been so rapid during the past three years that it has become 
necessary to rewrite a considerable portion of the book. The chapters 
on Fuels and Combustion, Engines, Turbines, Condensers, Finance and 
Economics have been entirely rewritten and the others have been 
revised and enlarged. The chapter on Finance and Economics con- 
tains many operating charts showing current practice in power plant 
cost accounting, typical load curves illustrating daily, monthfy and 
yearly load characteristics for a wide range in practice, and numerous 
tables giving recent results in cost of operation. Over three hundred 
new illustrations have been added and the entire text has been reset. 

Although primarily intended as a textbook for engineering students 
it is hoped that the new edition will be of interest to practicing 
engineers. 

G. F. G. 

Chicago, III., April, 1913. 



in 



CONTENTS 



PAGE 

CHAPTER I. — Elementary Steam Power Plants 1-10 

' 1. General 1 

2. Elementary Non-condensing Plant 2 

3. Non-condensing Plant; Exhaust Steam Heating 5 

4. Elementary Condensing Plant 7 

5. Condensing Plant with Full Complement of Heat-saving Devices. ... 10 

CHAPTER II. — Fuels and Combustion 11-85 

6. General 15 

7. Classification of Fuels 15 

8. Solid Fuels 15 

9. Composition of Coal 16 

10. Classification of Coals 16 

11. Anthracites 18 

12. Semi-anthracites 19 

13. Semi-bituminous 20 

14. Bituminous 20 

15. Lignite 23 

16. Peat or Turf 24 

17. Wood, Straw, Sawdust, Bagasse, Tanbark 25 

18. Combustion 29 

19. Calorific Value of Coals 32 

20. Air Required for Combustion 36 

21. Temperature Due to Combustion 43 

22. Heat Losses in Burning Coal 45 

23. Loss in Dry Chimney Gases 45 

24. Loss Due to Incomplete Combustion 47 

25. Loss of Fuel Through Grate 49 

26. Superheating Moisture in the Air 50 

27. Loss Due to Moisture in the Fuel 50 

28. Loss Due to the Presence of Hydrogen in the Fuel 51 

29. Loss Due to Visible Smoke 51 

30. Radiation and Minor Losses; Standby Losses 52 

31. Size of Coal — Bituminous 54 

32. Washed Coal 54 

33. Purchasing Coal 56 

34. Powdered Coal 58 

35. Furnaces for Burning Powdered Coal 59 

36. Types of Powdered Coal Burners 59 

37. Pinther Apparatus for Burning Powdered Coal 59 

38. Schwartzkopff Apparatus for Burning Powdered Coal 61 

39. Blake Apparatus for Burning Powdered Coal 62 

40. Triumph Apparatus for Burning Powdered Coal 63 

v 



VI CONTENTS 

PAGE 

41. Efficiency of Powdered-coal Furnaces 64 

42. Rate of Combustion with Powdered Coal 64 

43. Draft for Powdered Coal 65 

44. Storing Powdered Coal 65 

45. Depreciation of Powdered-coal Furnaces 65 

46. Cost of Pulverizing Coal 65 

47. Fuel Oil .' 66 

48. Chemical and Physical Properties of Fuel Oil 66 

49. Efficiency of Boiler with Fuel Oil 69 

50. Comparative Evaporative Economy of Oil and Coal 70 

51. Types of Oil Burners 70 

52. Furnaces for Burning Oil Fuel 75 

53. Atomization of Oil 79 

54. Oil-feeding Systems ; 79 

55. Oil Storage and Transportation 82 

56. Purchase of Fuel Oil 83 

57. Gaseous Fuels 85 

CHAPTER III. — Boilers 89-152 

58. General 89 

59. Classification 89 

60. Vertical Tubular Boilers 89 

61. Fire-box Boilers 91 

62. Fitzgibbons Boilers 91 

63. Scotch-Marine Boilers 93 

64. Robb-Mumford Boiler 95 

65. Horizontal Return Tubular Boiler 96 

66. Lyons Boiler 102 

67. Sederholm Boiler , 102 

68. Babcqck and Wilcox Boiler 104 

69. Heine Boiler 106 

70. Wickes Vertical Water-tube Boiler 107 

71. Parker Boiler 107 

72. Stirling Boiler 110 

73. The Bigelow-Hornsby Boiler 113 

74. Unit of Evaporation 114 

75. Heat Transmission 115 

76. Heating Surface 118 

77. Horse Power of a Boiler 120 

78. Grate Surface 121 

79. Boiler and Furnace Efficiency 123 

80. Boiler Performances 126 

81. Effect of Capacity on Efficiency 130 

82. Thickness of Fire 135 

83. Influence of Initial Temperature on Efficiency 138 

84. Cost of Boiler and Setting 141 

85. Selection of Type 141 

86. Grates 142 

87. Shaking Grates 144 

88. Blow-offs 145 



CONTENTS vii 

PAGE 

89. Dampers 146 

90. Water Gauges . 147 

91. Fusible Plugs 149 

92. Mechanical Tube Cleaners 150 

CHAPTER IV. — Smoke Prevention, Furnaces, Stokers 153 

93. General 153 

94. Hand-fired Furnaces 154 

' 95. Dutch Ovens 156 

96. Twin-fire Furnaces 157 

97. Chicago Settings for Hand-fired Return Tubular Boiler 159 

98. Wooley Smokeless Furnace 163 

99. Kent's Wing-wall Furnace 163 

100. Burke's Smokeless Furnace 164 

101. Down-draft Furnaces 164 

102. Steam Jets 167 

103. Parson Smokeless Furnace 168 

104. Heinrich Smokeless Furnace 169 

105. Luckenbach Smokeless Furnace 170 

106. Mechanical Stokers 170 

107. Chain Grates 172 

108. Step Grates, Front Feed 178 

109. Step Grates, Side Feed 180 

110. Jones Underfeed Stoker 182 

111. American Underfeed Stoker 183 

112. Taylor Underfeed Stoker 183 

113. Sprinkling Stokers 185 

114. Smoke Determinations 186 

115. Cost of Stokers 189 

CHAPTER V. — Superheated Steam; Superheaters 190-221 

116. General 190 

117. Economy of Superheat 191 

118. Limit of Superheat 192 

119. Properties of Superheated Steam 193 

120. Superheaters 194 

121. Babcock and Wilcox Superheater 200 

122. Stirling Superheater 201 

123. Foster Superheater 201 

124. Heine Superheater 205 

125. Independently Fired Superheaters 205 

126. Luckenbach Superheater 208 

127. Materials Used in Construction of Superheaters 209 

128. Extent of Superheating Surface 210 

129. Performance of Superheaters 216 

CHAPTER VI. — Coal and Ash-handling Systems 222-246 

130. General 222 

131. Coal Storage 222 

132. Coal Conveyors 224 



viii CONTENTS 

PAGE 

133. Hand Shoveling 224 

134. Bucket Conveyors 225 

135. Belt Conveyors 232 

136. Elevating Tower, Hand-car Distribution 233 

137. Overhead Storage, Bucket Hoist 235 

138. Elevating Tower, Cable-car Distribution 236 

139. Hoist and Trolley 236 

140. "Vacuum " Ash Conveyor 238 

141. Cost of Handling Coal and Ashes 242 

142. Coal Hoppers 242 

143. Coal Valves 245 

CHAPTER VII. — Chimneys 247-284 

144. Chimney Draft 247 

145. Chimney Formulas 252 

146. Height of Chimneys for Boilers Using Oil Fuel 259 

147. Classification of Chimneys 259 

148. Guyed Chimneys 260 

149. Self-sustaining Steel Chimneys 261 

150. Thickness of Plates 261 

151. Riveting . . 264 

152. Stability of Steel Chimneys 264 

153. Brick Chimneys 265 

154. Thickness of Walls 267 

155. Core and Lining 271 

156. Materials for Brick Chimneys 271 

157. Stability of Brick Chimneys 272 

158. Custodis Radial Brick Chimney 275 

159. Steel-Concrete Chimneys 275 

160. Breeching 280 

161. Chimney Foundations 281 

162. Chimney Efficiencies 282 

163. Cost of Chimneys 283 

CHAPTER VIII. — Mechanical Draft 285-306 

164. General 285 

165. Steam Jets 285 

166. Fan Draft 289 

167. Performance of Fans 292 

168. Determination of the Size of Fan 297 

169. Chimney vs. Mechanical Draft 300 

170. "Balanced Draft" 306 

CHAPTER IX. — Reciprocating Steam Engines 307-385 

171. Introductory 307 

172. The Ideal Engine 307 

173. The Carnot Cycle 308 

174. The Rankine Cycle 310 

175. Rankine Cycle with Incomplete Expansion 313 

176. Conventional Ideal Engine 317 



CONTENTS ix 

PAGE 

177. The Actual Engine 319 

178. Efficiency Standards 319 

179. Steam Consumption 320 

180. Heat Consumption 320 

181. Thermal Efficiency 321 

182. Mechanical Efficiency 323 

183. Efficiency Ratio 323 

184. Cylinder Efficiency 325 

185. Commercial Efficiency ". . 327 

186. Heat Losses in the Steam Engine 327 

187. Cylinder Condensation 327 

188. Leakage of Steam 329 

189. Clearance Volume 331 

190. Loss Due to Incomplete Expansion and Compression 332 

191. Loss Due to Wire Drawing 334 

192. Loss Due to Friction of the Mechanism 334 

193. Moisture 336 

194. Radiation and Minor Losses 336 

195. Heat Lost in the Exhaust 336 

196. Methods for Increasing Economy 337 

197. Effect of Increased Steam Pressure 337 

198. Receiver Reheaters; Intermediate Reheating 339 

199. Jackets 340 

200. Increasing Rotative Speed 341 

201. High-speed Single-valve Simple Engine 342 

202. High-speed Multi-valve Simple Engine 347 

203. Medium- and Low-speed Multi-valve Simple Engine , 349 

204. Compound Engines 350 

205. Triple and Quadruple Engines 359 

206. Effect of Condensing 361 

207. Throttling vs. Automatic Cut-off 364 

208. Influence of Superheat 367 

209. The Locomobile 376 

210. Uniflow or Straight-flow Engine 378 

211. Binary Vapor Engine 379 

212. Rotary Engines 382 

213. Cost of Engines 385 

CHAPTER X. — Steam Turbines 386-460 

214. Classification 386 

215. General Elementary Theory 388 

216. De Laval Turbine 390 

217. Elementary Theory 394 

218. Terry Turbine 405 

219. Elementary Theory; Terry Turbine 407 

220. Kerr Turbine 407 

221. De Laval Multi-stage Turbine 411 

222. Elementary Theory; Kerr Turbine 411 

223. Curtis Turbine 412 

224. Elementary Theory; Curtis Turbine 423 



x CONTENTS 



225. Westinghouse Impulse Turbine ^ 

226. Westinghouse Reaction Turbine; Parsons Singled Type.' .'.'.'"" 428 
til' ^tmghouse ^P^e-^action Turbine; Double-flow Type. 
228. Allis-Chalmers Turbine.. . . 



229. Elementary Theory; Parsons Turbine. . '.'.['. 4 q~ 

230. Low-pressure, Mixed-pressure and Bleeder Turbines 439 

231. Advantages of the Steam Turbine 44 6 

232. Efficiency and Economy of the Steam Turbine aao 

233. First Cost of Turbines. . t!? 

234. Cost of Operation f 6 

235. Influence of Superheat . . 

236. Influence of High Vacua 



455 



237. Tesla Bladeless Turbine. . fjt 

238. "Spiro" Turbine "" ™* 

459 

CHAPTER XL — Condensers , R1 ™ 

239. General * " " 46 

240. Function of the Condenser ..... ' .' .' ' ' ' ' ' ' ' ' ' ' ' ' 464 

241. Classification of Condensers 4fifi 

242. Standard Low-vacuum Jet Condensers 467 

243. Condensing Water, Jet Condenser 470 

244. Effect of Aqueous Vapor Upon the Degree of Vacuum at\ 

245. Injection Orifice ^ 

246. Volume of Condenser Chamber 474 

247. Injection and Discharge Pipes 474 

248. Siphon Condensers 47 . 

249. Size of Siphon Condensers 475 

250. Ejector Condenser 47fi 

251. Barometric Condenser 477 

252. Water-cooled Surface Condenser 48O 

253. High-vacuum Systems 484 

254. Cooling Water, Surface Condensers 491 

255. Extent of Water-cooling Surface 492 

256. Dry-air Surface Condenser 499 

257. Quantity of Air for Cooling [][ 500 

258. Saturated-air Surface Condenser 501 

259. Evaporative Surface Condenser 503 

260. Location and Arrangement of Condensers ' . 505 

261. Power Consumption of Condenser Auxiliaries 510 

262. Cost of Condensers r ^ 2 

263. Most Economical Vacuum e--, r 

264. Choice of Condensers ~ 16 

265. Water-cooling Systems. .. . " " 17 

266. Cooling Pond "" V.' 

267. Spray Pond .'.'.'.'.'.".'.'.'.'.'.'.'.'.'.'.'.'.".'.' 513 

268. Cooling Towers 520 

269. Parallel Comparison of Fan and Natural-draft Cooling Towers 523 

270. Hygrometry " 523 

271. Water-cooling Calculations " 52 q 

272. Test of Cooling Tower 535 



CONTENTS X1 



PAGE 



CHAPTER XII.— Feed-water Purifiers and Heaters ^'SI 

273. General • 543 

274. Chemical Purification ^ 

275. Boiler Compounds "V'^jmi ^±± 

276. Use of Kerosene and Petroleum Oils in Boiler Feed Water. 544 

277. Use of Zinc in Boilers 

278. Methods of Introducing Compounds 

279. Weights of Compounds Necessary ^ 

-280. Mechanical Purification 

281. Thermal Purification 547 

282. Purifying Plants 

283. Economy of Preheating Feed Water 

284. Classification of Feed-water Heaters ^ 

285. Open Heaters 557 

286. Open Heaters and Purifiers ■ ^ 

287. Temperatures in Open Heaters 

288. Pan Surface Required in Open Heaters 

289. Size of Shell, Open Heaters ^ Q 

290. Classification of Closed Heaters 



291. Closed Heaters, Water-tube. 



561 

292. Closed Heaters, Steam-tube 

293. Heating Surface, Closed Heaters • ^ 

294. Heat Transmission, Closed Heaters 

295. Open vs. Closed Heaters 

296. Through Heaters 572 

297. Induced Heaters 

298. Live-steam Heaters and Purifiers ^ 

299. Economizers ^ 

300. Value of Economizers • 

301. Factors Determining Installation of Economizers &<» 

302. Feed-water Temperature Due to Use of Economizers *>/» 

303. Choice of Feed-water Heating Systems 

CHAPTER XIII. — Pumps 58g 

304. Classification of Pumps 

305. Boiler-feed Pumps, Direct-acting Duplex ™> 

306. Boiler-feed Pumps, Direct-acting, Steam-actuated Gear MM 

307. Air and Vacuum Chambers 5Q5 

308. Water Pistons and Plungers ^ 

309. Performance of Piston Pumps ^ 

310. Size of Boiler-feed Pumps 6Q7 

311. Steam-pump Governors 6Q7 

312. Feed-water Regulators 

313. Power Pumps 612 

314. Injectors 612 

315. Positive Injectors -. 612 

316. Automatic Injectors 614 

317. Performance of Injectors 

318. Injector vs. Steam Pump as a Boiler Feed ^ 

319. Air Pumps 



xii CONTENTS 

PAGE 

320. Dean Wet-air Pump . ■ 618 

321. Size of Wet-air Pumps for Jet Condensers 619 

322. Edwards Air Pump 622 

323. Mullen Valveless Air Pump 622 

324. Wheeler "Rotrex" Air Pump 623 

325. Leblanc Air Pump 624 

326. Size of Wet-air Pumps for Surface Condensers 626 

327. Alberger Dry-air Pump ' 627 

328. Size of Dry-air Pumps for Surface Condensers 627 

329. Centrifugal Pumps 629 

330. Volute Centrifugal Pumps 630 

331. Turbine Centrifugal Pumps 630 

332. Field for Centrifugal Pumps 631 

333. Performance of Centrifugal Pumps 635 

334. Rotary Pumps 638 

335. Circulating Pumps 640 

336. Hot-well Pumps 641 

337. Air Lift 641 

CHAPTER XIV. — Sepakatoes, Traps, Drains 644-675 

338. Live-steam Separators 644 

339. Classification of Separators 645 

340. Reverse-current Steam Separators 646 

341. Centrifugal Steam Separators 647 

342. Baffle-plate Steam Separators 648 

343. Mesh Separators 648 

344. Location of Separators 649 

345. Exhaust-steam Separators and Oil Eliminators 650 

346. Exhaust Heads 653 

347. Drips 655 

348. Low-pressure Drips 655 

349. Size of Pipes for Low-pressure Drips 655 

350. High-pressure Drips 657 

351. Classification of Traps 657 

352. Float Traps 658 

353. Bucket Traps 659 

354. Dump or Bowl Traps 659 

355. Expansion Traps 660 

356. Differential Traps 663 

357. Location of Traps 664 

358. Drips Under Vacuum 666 

359. Drips Under Alternate Pressure and Vacuum 667 

360. The Steam Loop 668 

361. The HoUy Loop 669 

362. Returns Tank and Pump 671 

363. Office Building Drains , 672 

364. Radiation Drains 673 

| CHAPTER XV. — Piping and Pipe Fittings 676-749 

365. General 676 

366. Drawings 676 



CONTENTS xiii 

PAGE 

367. Materials for Pipes and Fittings 676 

368. Size and Strength of Commercial Pipe 679 

369. Screwed Fittings 680 

370. Flanged Fittings 682 

371. Pipe Coverings 692 

372. Expansion Due to Temperature Variation 694 

373. Pipe Supports and Anchors 697 

374. General Arrangements of High-pressure Steam Piping 698 

375. Main Steam Headers 707 

376. Flow of Steam in Pipes 710 

377. Equation of Pipes 719 

378. Friction Through Valves and Fittings 719 

379. Exhaust Piping, Condensing Plants 721 

380. Exhaust Piping, Non-condensing Plant, Webster Vacuum System. . . 721 

381. Exhaust Piping, Non-condensing Plant, Paul Heating System 725 

382. Automatic Temperature Control 726 

383. Feed-water Piping 727 

384. Flow of Water Through Orifices 730 

385. Stop Valves. 735 

386. Automatic Non-return Valves 738 

387. Emergency Valves 739 

388. Check Valves 740 

389. Blow-off Cocks and Valves 741 

390. Safety Valves 743 

391. Back-pressure and Atmospheric Relief Valves . 746 

392. Reducing Valves; Pressure Regulators 747 

393. Foot Valves 749 

CHAPTER XVI. — Lubricants and Lubrication 750-771 

394. General 750 

395. Vegetable Oils 750 

396. Animal Oils and Fats 750 

397. Mineral Oils 751 

398. Solid Lubricants 752 

399. Greases 753 

400. Qualification of Good Lubricants 754 

401. Identification of Oils 755 

402. Gravity 755 

403. Viscosity 756 

404. Flash Point 756 

405. Burning Point or Fire Test 756 

406. Acidity 756 

407. Cold Test 756 

408. Friction Test 756 

409. Atmospheric Surface Lubrication 757 

410. Intermittent Feed 757 

411. Restricted Feed 757 

412. Oil Bath 757 

413. Oil Cups 759 

414. Telescopic Oiler 759 



xiv CONTENTS 

PAGE 

415. Ring Oiler , 760 

416. Centrifugal Oiler 760 

417. Pendulum Oiler 761 

418. " Splash Oiling " 761 

419. Gravity Oil Feed 761 

420. Low-pressure Gravity System 761 

421. Compression Air Feed 762 

422. Cylinder Lubrication 764 

423. Cylinder Cups 764 

424. Hydrostatic Lubricator 765 

425. Forced-feed Cylinder Lubrication 766 

426. Central Systems 766 

427. Oil Filters 769 

CHAPTER XVII. — Testing and Measuring Apparatus 772-809 

428. General 772 

429. Weighing Fuel 772 

430. Measuring of Feed Water 774 

431. Actual Weighing of Feed Water 774 

432. Worthington Weight Determinator , 774 

433. Kennicott Water Weigher 775 

434. Wilcox Water Weigher 776 

435. Weir Measuring Devices 777 

436. Pressure Water Meters 777 

437. Venturi Meter 778 

438. Orifice Measurements 780 

439. Measurement of Steam 781 

440. Weighing Condensed Steam 781 

441. Steam Meters 781 

442. "Gebhardt" Steam Meter 783 

443. General Electric Steam Meters 785 

444. St. John's Steam Meter 790 

445. Pressure Gauges 792 

446. Measurement of Temperature 792 

447. Wilsey Relative Efficiency Indicator 797 

448. Power Measurements 798 

449. Measurement of Speed 798 

450. Steam-engine Indicators 799 

451. Dynamometers 800 

452. Flue Gas Analysis — The Orsat Apparatus 801 

453. Williams Improved Gas Apparatus 802 

454. Little Modified Orsat Apparatus 803 

455. Simmance-Abady C0 2 Recorder 804 

456. Uehling Gas Calorimeter 805 

457. Moisture in Steam 806 

458. Fuel Calorimeters 808 

CHAPTER XVIII. — Finance and Economics. — Cost of Power . . 810-860 

459. General Records 810 

460. Permanent Statistics 812 



CONTENTS xv 

PAGE 

461. Operating Records 812 

462. Output and Load Factor 815 

463. Cost of Operation 820 

464. Fixed Charges 820 

465. Interest 820 

466. Depreciation 823 

467. Maintenance 829 

468. Taxes and Insurance 829 

469. Operating Costs 830 

470. Labor, Attendance, Wages 830 

471. Cost of Fuel 833 

472. Oil, Waste and Supplies 834 

473. Repairs and Maintenance 834 

474. Cost of Power 834 

475. Elements of Power Plant Design 853 

CHAPTER XIX. — Typical Specifications 861-884 

476. Specifications for a Cross- compound Non-condensing Engine 861 

477. Specifications for a Return Tubular Boiler 865 

478. Specifications for a Condenser Plant 869 

479. Specifications for a Piping System 871 

480. Government Specifications for Purchasing Coal ■ 880 

CHAPTER XX. — Typical Steam Turbine Stations 885-898 

Steam Turbine Stations; Commonwealth Edison Co., Chicago. 

CHAPTER XXI. — Typical Modern Isolated Station 899-911 

A Typical Modern Isolated Station; W. H. McElwain Company, Man- 
chester, N. H. 

APPENDICES A-H 

A. S. M. E. Codes of 1912. Preliminary Report of Committee on Power 
Tests. 

APPENDIX A. — General Instructions 912 

APPENDIX B. — Rules for Conducting Evaporative Tests of Boilers 916 

APPENDIX C. — Rules for Conducting Tests of Reciprocating En- 
gines 927 

APPENDIX D. — Rules for Conducting Tests of Steam Turbines 

and Turbo-Generators 939 

APPENDIX E. — Rules for Conducting Tests of Complete Power 

Plants 943 

APPENDIX F. — Rules for Conducting Duty Trials of Steam Pump- 
ing Machinery 947 

APPENDIX G. — Rules for Conducting Tests of Complete Steam- 
pumping Machinery 953 



xvi CONTENTS 

PAGE 



APPENDIX H. — Rules for Conducting Tests of Steam-driven Com- 
pressors, Blowers and Fans 956 

APPENDIX I. — Properties of Saturated Steam 961 

APPENDIX J. — Equivalent Values of Mechanical and Electrical 

Units 964 

APPENDIX K. Miscellaneous Conversion Tables 965 

APPENDIX L. — Temperature-Entropy Diagram (Mollier Diagram) 966 



STEAM POWER PLANT 
ENGINEERING. 



CHAPTER I. 

ELEMENTARY STEAM POWER PLANTS. 

1. General. — The fundamental object of any power plant is the 
conversion of energy from one form into another at the least ultimate 
cost. This involves not only the cost of converting the energy into the 
desired form, but also the cost of distribution and application. 

The most efficient plant, thermally, in the conversion of energy 
from one form to another, is not necessarily the most economical com- 
mercially, since the various items involved in effecting this conversion 
may more than offset the gain over a less efficient plant. There is no 
question as to the low cost of power generated by hydro-electric plants, 
but when the cost of transmission and the overhead charges are taken 
into consideration the economy is not so evident and may be com- 
pletely neutralized. From a purely thermal standpoint, and as a 
means of conserving our natural resources, the producer-gas-electric 
plant is vastly superior to the ordinary steam-electric plant for power 
purposes, but the fuel item is only one of the many involved in the 
total cost. It is the commercial efficiency which enables the steam 
power plant, with its extravagant waste of fuel, to compete successfully 
with the gas producer, internal-combustion engine and hydro-electric 
plant. 

A station which distributes power to a number of consumers more 
or less distant, is called a Central Station. When the distances are 
very great, electrical current of high tension is frequently employed, 
and is transformed and distributed at convenient points through Sub- 
stations. A plant designed to furnish power or heat to a building or a 
group of buildings under one management is called an Isolated Station. 
For example, the power plant of an office building is usually called an 
isolated station. 

When the exhaust steam from the engines is discharged at approxi- 
mately atmospheric pressure the plant is said to be operating non- 

1 



STEAM POWER PLANT ENGINEERING 



condensing. When the exhaust steam is condensed, reducing the back 
pressure on the piston by the partial vacuum thus formed, the plant 
is said to operate condensing. 

When the exhaust steam may be used for manufacturing, heating, or 
other useful purposes, as is frequently the case in various manufac- 
turing establishments, and in large office buildings, it is usually more 
economical to run non-condensing, while power plants for electric 
lighting and power, pumping, stations, air-compressor plants, and others, 
in which the load is fairly constant and the exhaust steam is not re- 
quired for heating, are generally operated condensing. 



Water 
Gauge 




Injector 



Fig. 1. Elementary Non-condensing Plant. 

2. Elementary Non-condensing Plant. — Fig. 1 gives a diagrammatic 
outline of the essential elements of the simplest form of steam power 
plant. The equipment is complete in every respect and embodies all 
the accessories necessary for successful operation. Wliere a small 
amount of power is desired at intermittent periods, as in hoisting 
systems, threshing outfits and traction machinery, the arrangement is 
substantially as illustrated. The output in these cases seldom exceeds 
50 horse power and the time of operation is usually short, so the cheapest 
of appliances are installed, simplicity and low first cost being more im- 
portant than economy of fuel. 



ELEMENTARY STEAM POWER PLANTS 3 

Such a plant has three essential elements: (1) The furnace, (2) the 
boiler, and (3) the engine. Fuel is fed into the furnace, where it is 
burned. A portion of the heat liberated from the fuel by combustion 
is absorbed by the water in the boiler, converting it into steam under 
pressure. The steam being admitted to the cylinder of the engine does 
work upon the piston and is then exhausted through a suitable pipe 
to the atmosphere. The process is a continuous one, fuel and water 
being fed into the furnace and the boiler in proportion to the power 
demanded. 

In such an elementary plant, certain accessories are necessary for 
successful operation. The grate for supporting the fuel during com- 
bustion consists of a cast-iron grid or of a number of cast-iron bars 
spaced in such a manner as to permit the passage of air through the 
fuel from below. The solid waste products fall through or are " sliced" 
through the grate bars into the ash pit, from which they may be re- 
moved through the ash door. The latter acts also as a means of regu- 
lating the supply of air below the grate. Fuel is fed into the furnace 
through the fire door, and when occasion demands, air may be supplied 
above the bed of fuel by means of this door. The combustion chamber 
is the space between the bed of fuel and the boiler heating surface, its 
office being to afford a space for the oxidation of the combustible gases 
from the solid fuel before they are cooled below ignition temperature by 
the comparatively cool surfaces of the boiler. The chimney or stack 
discharges the products of combustion into the atmosphere and serves 
to create the draft necessary to draw the air through the bed of fuel. 
Various forced-draft appliances are sometimes used to assist or to en- 
tirely replace the chimney. The heating surface is that portion of the 
boiler area which comes into contact with the hot furnace gases, absorbs 
the heat and transmits it to the water. In the small plant illustrated 
in Fig. 1, the major portion of the heating surface is composed of a 
number of fire tubes below the water line, through which the heated 
gases pass. The superheating surface is that portion of the heating 
surface which is in contact with the heated gases of combustion on one 
side and steam on the other. The volume above the water level is 
called the steam space. Water is forced into the boilers either by a 
feed pump or an injector. In small plants of the type considered, steam 
pumps are seldom employed; the injector answers the purpose and is 
considerably cheaper. A safety valve connected to the steam space of 
the boiler automatically permits steam to escape to the atmosphere if 
an excessive pressure is reached. The water level is indicated by try 
cocks or by a gauge glass, the top of which is connected with the steam 
space and the bottom with the water space. Try cocks are small valves 



4 STEAM POWER PLANT ENGINEERING 

placed in the water column or boiler shell, one at normal water level, 
one above it, and one below. By opening the valves from time to 
time the water level is approximately ascertained. They are ordinarily 
used in case of accident to the gauge glass. Fusible plugs are frequently 
inserted in the boiler shell at the lowest permissible water level. They 
are composed of an alloy having a low fusing point which melts when 
in contact with steam, thus giving warning by the blast of the escaping 
steam if the water level gets dangerously low. The blow-off cock is a 
valve fitted to the lowest part of the boiler to drain it of water or to 
discharge the sediment which deposits in the bottom. The steam out- 
let of a boiler is usually called the steam nozzle. 

The essential accessories of the simple steam engine include: A 
throttle valve for controlling the supply of steam to the engine; the 
governor, which regulates the speed of the engine by governing the 
steam supply; the lubricator, attached to the steam pipe, which is 
usually of the " sight-feed" class and provides for lubrication of piston 
and valve. Lubrication of the various bearings is effected by oil cups 
suitably located. Drips are placed wherever a water pocket is apt to 
form in order that the condensation may be drained. The apparatus 
to be driven by the engine may be direct connected to the crank shaft 
or belted to the flywheel or geared. 

In small plants of this type no attempt is made to utilize the exhaust 
steam except in instances where the stack is too short to create the 
necessary draft, in which case the exhaust may be discharged up the. 
stack. If the draft is produced by convection of the heated gases in 
the chimney, the fuel is said to be burned under natural draft; if the 
natural draft is assisted by the exhaust steam, the fuel is said to be 
burned under forced draft. The power realized from a given weight of 
fuel is very low and seldom exceeds 2\ per cent of the heat value of the 
fuel. The distribution of the various losses in a plant of, say, 40 
horse power is approximately as follows : 

b.t.u. 

Heat value of 1 pound of coal 14,500 

Boiler and furnace losses, 50 per cent 7,250 

Heat of the steam, 50 per cent 7,250 

Heat equivalent of one horse power hour 2,545 

Heat used to develop one horse power hour (50 pounds steam per 

horse power hour, pressure 80 pounds gauge, feed water 62 degrees F.) 57,500 

Per cent. 

2 545 
Percentage of heat in the steam, realized as work, ^^qq • • • • 4 - 4 

2 545 
Percentage of heat value of the coal realized as work, _ eqq _ q 5Q ^'^ 



ELEMENTARY STEAM POWER PLANTS 5 

In Europe small non-condensing plants are developed to a high 
degree of efficiency. Through the use of highly superheated steam, 
specially designed engines and boilers, plants of this type as small as 
40 horse power are operated with over-all efficiencies of from 10 to 12 
per cent. 

The power plant of the modern locomotive is very much like that 
illustrated in Fig. 1, the main difference lying in the type of boiler and 
engine. The entire exhaust from the engine is discharged up the 
stack through a suitable nozzle, since the extreme rate of combustion 
requires an intense draft. The engine is a highly efficient one com- 
pared with that in the illustration, and the performance of the boiler 
is more economical. In average locomotive practice about 6 per cent 
of the heat value of the fuel is converted into mechanical energy at 
the draw bar. In general, a non-condensing steam plant in which the 
heat of the exhaust is wasted is very uneconomical of fuel, even 
under the most favorable conditions, and seldom transforms as much 
as 7 per cent of the heat value of the fuel into mechanical energy. 

3. Non-condensing Plant. Exhaust Steam Heating. — Fig. 2 gives 
a diagrammatic arrangement of a simple non-condensing plant differ- 
ing from Fig. 1 in that the exhaust steam is used for heating pur- 
poses. This shows the essential elements and accessories, but omits 
a number of small valves, by-passes, drains, and the like for the sake 
of simplicity. The plant is assumed to be of sufficient size to warrant 
the installation of efficient appliances. Steam is led from the boiler 
to the engine by the steam main. The moisture is removed from the 
steam before it enters the cylinder by a steam separator. The moisture 
drained from the separator is either discharged to waste or returned to 
the boiler. The exhaust steam from the engine is discharged into the 
exhaust main where it mingles with the steam exhausted from the steam 
pumps. Since the exhaust from engines and pumps contains a large 
portion of the cylinder oil introduced into the live steam for lubricat- 
ing purposes, it passes through an oil separator before entering the 
heating system. After leaving the oil separator the exhaust steam is 
diverted into two paths, part of it entering the feed-water heater where 
it condenses and gives up heat to the feed water, and the remainder 
flowing to the heating system. During warm weather the engine 
generally exhausts more steam than is necessary for heating purposes, 
in which case the surplus steam is automatically discharged to the 
exhaust head through the back-pressure valve. The back-pressure valve 
is, virtually, a large weighted check valve which remains closed when 
the pressure in the heating system is below a certain prescribed amount, 
but which opens automatically when the pressure is greater than this 



STEAM POWER PLANT ENGINEERING 




ELEMENTARY STEAM POWER PLANTS 7 

amount. During cold weather it often happens that the engine ex- 
haust is insufficient to supply the heating system, the radiators con- 
densing the steam more rapidly than it can be supplied. In this case 
live steam from the boiler is automatically fed into the main heating 
supply pipe through the reducing valve. 

The condensed steam and the entrained air which is always present 
are automatically discharged from the radiators by a thermostatic valve 
into the returns header. The thermostatic valve is so constructed that 
when in contact with the comparatively cool water of condensation it 
remains open and when in contact with steam it closes. The vacuum 
pump or vapor pump exhausts the condensed steam and air from the 
returns header and discharges them to the returns tank. The small 
pipe S admits cold water to the vacuum pump and serves to condense 
the heated vapor, and at the same time supply the necessary make-up 
water to the system. The returns tank is open to the atmosphere so 
that the air discharged from the vacuum pump may escape. From 
the returns tank the condensed steam gravitates to the feed-water 
heater where its temperature is raised to practically that of the exhaust 
steam. The feed water gravitates to the feed pump and is forced into 
the boiler. There are several systems of exhaust steam heating in 
current practice which differ considerably in details, but, in a broad 
sense, are similar to the one just described. The more important of 
these will be described later on. 

During the summer months when the heating system is shut down, 
the plant operates as a simple non-condensing station and practically 
all of the exhaust steam, amounting to perhaps 60 per cent of the heat 
value of the fuel, is wasted. The total coal consumption, therefore, is 
charged against the power developed. During the winter months, 
however, all, or nearly all, of the exhaust steam may be used for heating 
purposes and the power becomes a relatively small percentage of the 
total fuel energy utilized. The percentage of heat value of the fuel 
chargeable to power depends upon the size of the plant, the number 
and character of engines and boilers, and the conditions of operation. 
It ranges anywhere from 50 to 100 per cent for the summer months 
and may run as low as 6 per cent for the winter months. This is on the 
assumption, of course, that the engine is debited only with the differ- 
ence between the coal necessary to produce the heat entering the cyl- 
inder and that utilized in the heating system. 

4. Elementary Condensing Plant. — Under the most favorable con- 
ditions a non-condensing plant can never be expected to realize more 
than 7 per cent of the heat value of the fuel as power. In large non- 
condensing power stations the demand for exhaust steam is usually 



STEAM POWER PLANT ENGINEERING 












^ 




i\\\«A-A > SSSSSSSS^SSSSSS < ^S^L" 



^ 



===3sr 5 vssssss sss s s ss ^^^^V^k^^^ 



i 






ELEMENTARY STEAM POWER PLANTS 9 

limited to the heating of the feed water, and as only 12 or 15 per 
cent can be utilized in this manner, the greater portion of the heat 
in the exhaust is lost. Non-condensing engines require from 20 to 
60 pounds of steam per hour for each horse power developed. On the 
other hand in condensing engines the steam consumption may be re- 
duced to as low as 10 pounds per horse-power hour. The saving of 
fuel is «at once apparent. 

Fig. 3 gives a diagrammatic arrangement of a simple condensing 
plant in which the back pressure on the engine is reduced by condens- 
ing the exhaust steam. A different type of boiler from that in Fig. 1 
or Fig. 2 has been selected, for the purpose of bringing out a few of the 
characteristic elements. The products of combustion instead of pass- 
ing directly through fire tubes to the stack as in Fig. 1 are deflected 
back and forth across a number of water tubes, by the bridge wall and a 
series of baffles. After imparting the greater part of their heat to the 
heating surface the products of combustion escape to the chimney 
through the breeching or flue. The rate of flow is regulated by a damper 
placed in the breeching as indicated. 

The steam generated in the boiler is led to the engine through the 
main header. The steam is exhausted into a condenser in which its 
latent heat is absorbed by injection or cooling water. The process 
condenses the steam and creates a partial vacuum. The condensed 
steam, injection water, and the air which is invariably present are 
withdrawn by an air pump and discharged to the hot well. In case the 
vacuum should fail, as by stoppage of the air pump, the exhaust steam 
is automatically discharged to the exhaust head by the atmospheric 
relief valve, and the engine will operate non-condensing. The atmos- 
pheric relief valve is a large check valve which is held closed by atmos- 
pheric pressure as long as there is a vacuum in the condenser. When 
the vacuum fails the pressure of the exhaust becomes greater than 
that of the atmosphere and the valve opens. 

The feed water may be taken from the hot well or from any other 
source of supply and forced into the heater. In this particular case it is 
taken from a cold supply and upon entering the heater is heated by the 
exhaust steam from the air and feed pumps. From the heater it gravi- 
tates to the feed pump and is forced into the boiler. Various other 
combinations of heaters, pumps, and condensers are necessary in many 
cases, depending upon the conditions of operation. Feed pumps, air 
pumps, and in fact all small engines used in connection with a steam 
power plant are usually called auxiliaries. 

A well-designed station similar to the one illustrated in Fig. 3 is 
capable of converting about 10 per cent of the heat value of the fuel 



10 STEAM POWER PLANT ENGINEERING 

into mechanical energy. The various heat losses are approximately 
as follows : 

BOILER LOSSES. Per Cent. 

Loss due to fuel falling through the grate 2 

Loss due to incomplete combustion 2 

Loss to heat carried away in chimney gases 23 

Radiation and other losses 8 

Total . 35 

Heat used by engines and auxiliaries (16 pounds of steam per B.t.u. 

i.h.p.-hour, pressure 150 pounds, feed water 210° F.) . . . 16,250 

Engine and generator friction, 5 per cent 812 

Leakage, radiation, etc., 2 per cent 325 

Total 17,387 

Heat equivalent of one electrical horse power 2,545 

Percentage of the heat value of the steam converted into elec- Per Cent. 

trical energy 14.7 

Percentage of heat value of fuel converted into electrical energy 

2545 X 0.65 

17,387 95 

In large central stations equipped with turbo-generators using super- 
heated steam, an over-all efficiency from switchboard to coal pile of 
12 per cent is not unusual with a maximum of about 14 per cent. 

5. Condensing Plant with Full Complement of Heat-saving Appli- 
ances. — When fuel is costly it frequently becomes necessary for the 
sake of economy to reduce the heat wastes as much as possible. The 
chimney gases, which in average practice are discharged at a tem- 
perature between 450 and 550 degrees F., represent a loss of 20 to 
30 per cent of the total value of the fuel. If part of the heat could 
be reclaimed without impairing the draft the gain would be directly 
proportional to the reduction in temperature of the gases. Again, in 
some types of condensers all of the steam exhausted by the engine 
is condensed by the circulating water and discharged to waste. If 
provision could be made for utilizing part of the exhaust steam for 
feed-water heating, the efficiency of the plant could be correspondingly 
increased. In many cases the cost of installing such heat-saving de- 
vices would more than offset the gain effected, but occasions arise where 
they give marked economy. 

Fig. 4 gives a diagrammatic arrangement of a condensing plant in 
which a number of heat-reclaiming devices are installed. The plant is 
assumed to consist of a number of engines, boilers, and auxiliaries. 
Coal is automatically transferred from the cars to coal hoppers placed 
above the boiler, by a system of buckets and conveyors. These hoppers 
store the coal in sufficient quantities to keep the boiler in continuous 



ELEMENTARY STEAM POWER PLANTS 



11 




12 STEAM POWER PLANT ENGINEERING 

operation for some time. From the hoppers the coal is fed inter- 
mittently to the stoker by means of a down spout. The stoker feeds the 
furnace in proportion to the power demanded and automatically re- 
jects the ash and refuse to the ash pit. The ashes are removed from 
the ash pit, when occasion demands, and are transferred to the ash 
hopper by the same system of buckets and conveyor which handles 
the coal. The ash hopper is usually placed alongside the coal hoppers 
and is not unlike them in general appearance and construction. 

The products of combustion are discharged to the stack through the 
flue or breeching. Within the flue is placed a feed-water heater called 
an economizer, the function of which is to absorb part of the heat from 
the gases on their way to the chimney. The heat reclaimed by the 
economizer varies widely with the conditions of operation and ranges 
between 5 and 20 per cent. Since the economizer acts as a resistance 
to the passage of the products of combustion it is sometimes necessary 
to increase the draft either by increasing the height of the chimney 
or, as is the usual practice, by using a forced-draft system. 

Part of the heat of the exhaust steam is reclaimed by a vacuum heater 
which is placed in the exhaust line between engine and condenser. 
For example, if the feed water has a normal temperature of 60 degrees F. 
and the vacuum in the condenser is 26 inches, the vacuum heater will 
raise the temperature of the feed to, say, 120 degrees F., thereby effect- 
ing a gain in heat of approximately 6 per cent. If the feed supply is 
taken from the hot well the vacuum heater is without purpose, as the 
temperature of the hot well will not be far from 120 degrees F. 

Referring to the diagram, the path of the steam is as follows: From 
the boiler it flows through the boiler lead to the main header or equaliz- 
ing pipe. From the main header it flows through the engine lead to the 
high-pressure cylinder. The exhaust steam discharges from the low- 
pressure cylinder through the vacuum heater and into the condenser. 
Part of the exhaust steam is condensed in the vacuum heater and gives 
up its latent heat to the feed water. The remainder is condensed by 
the injection water which is forced into the condenser chamber by the 
circulating pump. The condensed steam and circulating water gravitate 
through the tail pipe to the hot well. The air which enters the con- 
denser either as leakage or entrainment is withdrawn by the air pump. 
The steam exhausted by the feed pump, air pump, stoker engine, and 
other steam-driven auxiliaries is usually discharged into the atmospheric 
heater, which still further heats the feed water. 

Referring to the feed water, the circuit is as follows: The pump 
draws in cold water at a temperature of, say, 60 degrees F., and forces 
it in turn through the vacuum heater, the atmospheric heater, and the 



ELEMENTARY STEAM POWER PLANTS 13 

economizer into the boiler. The vacuum heater raises the temperature 
to 120 degrees F., the atmospheric heater increases it to 212 degrees F., 
and the economizer still further to about 300 degrees F. The heat 
reclaimed by this series of heaters is evidently the equivalent of that 
necessary to raise the feed water from 60 degrees F. to 300 degrees F., 
or approximately 24 per cent of the total steam supplied. In some 
plants- the economizer only is installed; in others the economizer and 
atmospheric heater are deemed desirable; still others utilize all three. 
The distribution of the heat losses in a plant of this type using satu- 
rated steam and operating under favorable conditions is approximately 
as follows: 

PerCent. B.t.u. 

Delivered to engine, 15 pounds steam per i.h.p.-hour; 

pressure 150 pounds, feed 60° F 100 17,482 

Delivered to feed pump 1.5 262 

Delivered to circulating pump 1.5 262 

Delivered to air pump 2 349 1 

Delivered to small auxiliaries 1.5 262 

Loss in leakage and drips 0.5 87 

Engine and generator friction 5 874 

Radiation and minor losses 1 175 

Total 19,753 

Per Cent. B.t.u. 

Returned by vacuum heater 5.5 1,086 

Returned by atmospheric heater 7.9 1,560 

Returned by economizer 9.7 1,916 

Total 23.1 4,562 

Net heat delivered to engine in the form of steam to pro- 
duce one electrical horse power, 19,753 — 4,562 . . . . 15,191 

2 545 

Percentage converted to electrical power ' .... 16.7 

Boiler efficiency ' 70 

Percentage of heat value of fuel necessary to produce one 

electrical horse power at switchboard — ft ' — . . 11.7 

The preceding figures give the results of very good practice. So 
much depends upon the size and character of the prime movers, the 
nature of the fuel, and the conditions of operation that no definite 
figure can be given for the percentage of heat converted to power in a 
given type of station. Six per cent represents good average practice 
in a non-condensing plant and 10 per cent in a condensing plant using 
saturated steam. Pumping stations operating continuously under full 
load have realized as much as 15 per cent of the total heat value of the 
fuel, but such performances are practically unobtainable in connection 



14 STEAM POWER PLANT ENGINEERING 

with steam-driven electrical power plants with the usual peak loads. 
Steam power plants as a class are very wasteful of fuel at the best. 

One of the best recorded performances to date (March, 1912) of a 
steam-electric power plant is that of the Pacific Light and Power 
Company at Redondo, Cal. When operating under regular commer- 
cial conditions approximately 14 per cent of the available heat of the 
fuel (crude oil) is realized as power at the switchboard. This includes 
all standby losses. For a detailed description of the plant and the 
results of the acceptance tests, see Jour, of Elec. Gas and Power, Aug. 
22, 1908. 

In Europe a combined plant efficiency of 15 per cent is not uncom- 
mon. Even small semi-portable plants of 40 to 200 horse power are 
operated with over-all efficiencies as high as 14 per cent. In these 
small plants the engine, boiler, and auxiliaries are combined, permit- 
ting a high degree of superheat with minimum heat losses. A 40- 
horse-power plant tested by Professor Josse of the Royal Technical 
School, Germany, gave the following results: coal consumed per brake 
h.p.-hour, 1.23 pounds, corresponding to an over-all efficiency of 14.2 
per cent. Steam consumption, 9.5 pounds per i.h.p.-hour. Boiler and 
superheated efficiency, 77.7 per cent. (See Zeit. des Ver. Deut. Ing., 
March 18 and 25, 1911, and Power, Sept. 27, 1910, p. 1714. See Fig. 
226.) 



CHAPTER II. 

FUELS AND COMBUSTION. 

6. General. — The subject of fuels and combustion has been so 
extensively treated by various authorities that a comprehensive dis- 
cussion would be without purpose here, but in order to bring out more 
clearly the matter pertaining to the commercial design and operation 
of steam power plants a few of the essential elements will be briefly 
treated. 

The fuels used for steam making are coal, coke, wood, peat, mineral 
oil, natural and artificial gases, refuse products such as straw, manure, 
sawdust, tan bark, bagasse, and occasionally corn and molasses. 

In most cases that fuel is selected which develops the required power 
at the lowest cost, taking into consideration all of the circumstances 
that may affect its use. Occasionally the disposition of waste products 
is a factor in the choice, but such instances are uncommon. The 
boilers and furnaces are designed to suit the fuel selected. 

7. Classification of Fuels. — Fuels may be divided into three classs 
as follows: 

1. Solid fuels. 

a. Natural: straw, wood, peat, coal. 

b. Prepared: charcoal, coke, peat, and other briquettes. 

2. Liquid fuels. 

a. Natural: crude oils. 

b. Prepared: distilled oils, alcohol, molasses. 

3. Gaseous fuels. 

a. Natural: natural gas. 

b. Prepared: coal gas, water gas, producer gas, oil gas. 

8. Solid Fuels. — Solid fuels are of vegetable origin and exist in a 
variety of forms between that of a comparatively recent cellulose growth 
and that of nearly pure carbon as anthracite coal. They owe their 
forms to the conditions under which they were created or to the geo- 
logical changes which they have undergone. With each succeeding 
stage the percentage of carbon increases. The chemical changes are 
approximately as follows: 

15 



16 



STEAM POWER PLANT ENGINEERING 



Substance. 


Carbon. 


Hydrogen. 


Oxygen. 


Pure cellulose 

Wood 


Per Cent 
44.44 
52.65 
59.57 
66.04 
73.18 
75.06 
89.29 
91.58 

100.00 


Per Cent 
6.17 
5.25 
5.96 
5.27 
5.58 
5.84 
5.05 
3.96 


Per Cent 
49.39 
42 10 


Peat 


34 47 


Lignite . . . 


28 69 


Brown coal 

Bituminous coal 


21.14 
19 10 


Semi-bituminous coal 

Anthracite 

Graphite 


6.66 
4.46 



All natural solid fuels contain more or less earthy or inorganic matter 
which is not combustible and therefore remains as ash, while the 
organic matter is consumed. Sometimes the percentage of ash is so 
great as to render them valueless for steam-making purposes. 

Origin and Formation of Fuel: Engng, Aug. 23, 1901; Am. Geol., Feb., 1899; 
Col. Guard, Sept. 10, 1897, Oct. 1, 1897, Jan. 14, 1898, Jan. 28, 1898, March 18, 
1898, Sept. 14, 1900; Ec. Geol., Oct., 1905; Eng. U.S., April 1, 1903; Ir. and Coal 
Td. Review, Feb. 4, 1898, July 13, 1906. 

9. Composition of Coal. — The uncombined carbon in coal is known 
as fixed carbon, while the hydrocarbons and other gaseous compounds 
which distill off on application of heat constitute the volatile matter. 
Refractory earths and moisture are found in varying quantities in 
different classes of coal and as they are incombustible tend to reduce 
the heat value of the fuel. That part of the fuel which is dry and free 
from ash is called the combustible, though the nitrogen and oxygen in 
the volatile matter are not actually combustible. The term "pure 
coal" has been suggested in this connection and is meeting with much 
favor. (Jour. W.S.E. 11-757.) The various elementary constituents 
of a fuel must be determined by a careful chemical analysis, but in 
most cases it is only necessary to know the heating value, the per cent 
of moisture and ash, and perhaps the per cent of sulphur. Tables 1 to 
4 show the composition of a number of American coals and give a 
good idea of their chemical and physical characteristics. 

10. Classification of Coals. — Coals and allied substances have been 
variously classified according to 

1. Oxygen-hydrogen ratio, or Gruner's classification. 

2. Fixed carbon and volatile combustible matter. 

3. Fuel ratio, or the ratio of the fixed carbon to the volatile com- 
bustible matter. 

4. Calorific value. 

5. Fixed carbon. 



FUELS AND COMBUSTION 



17 



6. Total carbon. 

7. Hydrogen. 

8. Carbon-hydrogen ratio, or the ratio of the total carbon to the 
hydrogen. 

Gruner's classification is as follows: 

(Eng. and Min. Jour., July 25, 1874.) 





Ratio ■== • 
ti 




Ratio ==■ 
U 


Anthracite 


1 to 0.75 
4tol 
5 


Peat 


6 to 5 


Bituminous 


Wood 


7 


Lignite 


Cellulose 


8 



Kent's classification, according to the constituents of the combustible 
is as follows (Steam Boiler Practice) : 



Anthracite. 

Semi-anthracite 

Semi-bituminous 

Bituminous — Eastern . 
Bituminous — Western. 
Lignite 



Per Cent of Dry Combustible. 



Fixed Carbon. 


Volatile Matter. 


97 to 92.5 


3 to 7.5 


92.5 to 87.5 


7.5 to 12.5 


87.5 to 75 


12.5 to 25 


75 to 60 


25 to 40 


65 to 50 


35 to 50 


Under 50 


Over 50 



Gruner's, Kent's, and the other schemes of classification outlined 
above, with the exception of the carbon-hydrogen ratio, are more or 
less unsatisfactory, since the groups are not as clearly defined as indi- 
cated and overlap to a considerable extent. 

The U. S. Geological Survey proposes the following classification 
according to the carbon-hydrogen ratio which appears to apply satis- 
factorily to all grades of coal. 

(Compiled from Report of Government Coal Testing Plant, Professional Paper No. 48, 1906.) 



Group. 


Class. 


Example. 


Carbon-hydrogen 
Ratio. 


A 


Graphite 






B 


Anthracite 

Anthracite. . . 


*Buck Mountain, Pa 


to 30 


C 


*Scranton, Pa. 


30 to 26 


D 


Semi-anthracite. . . . 
Semi-bituminous . . . 

Bituminous 

...do 


*Bernice Basin, Pa. . . . 


26 to 23 


E. . 


Spadra Bed, Ark 


23 to 20 


F 


New River, W. Va. . . 


20 to 17 


G 


Connelsville Field, Pa 

Marion County, 111 

Red Lodge, Mont 


17 to 14.4 


H. ... 


...do 


14.4 to 12.5 


I .... 


. .do 


12.5 to 11.2 


J 


Lignite 

Peat 


Gallup Field, N. M 


11.2 to 9.3 


K. . 




9.3 to 


L... 


Wood 




7.2 











* Not included in Government's Report. 



18 



STEAM POWER PLANT ENGINEERING 



For a classification of various coals according to the calorific value, 
fixed carbon, total carbon and hydrogen content, consult government 
report. 

Classification of Coals: Prac. Engr. U. S., Jan., 1910; Mines and Minerals, 
Feb., 1911; Am. Inst. Min. Engrs., May, 1906, Sept., 1905; Eng. Mag., Jan., 1912. 

11. Anthracites. — These are the most perfect coals and consist 
almost entirely of carbon; they contain very little hydrocarbon and 
burn with little or no smoke, are slow to ignite, burn slowly, and break 
into small pieces when rapidly heated. They require a very large 
grate of about twice the surface necessary for bituminous coal. Large 
sizes may be burned in almost any kind of a furnace and with moderate 
draft. 

TABLE 1. 

COMPOSITION OF TYPICAL AMERICAN ANTHRACITE COALS. 



Proximate analysis: 

Water 

Volatile matter . . 

Fixed carbon 

Ash 



Ultimate analysis: 

Carbon 

Hydrogen 

Nitrogen 

Oxygen 

Sulphur 

Ash 



Calorific value: 
Calorimeter .... 
Dulong's formula. 



Classification: 
Carbon-hydrogen ratio 
Fuel ratio 



0.84 

6.67 

85.66 

6.83 



100.00 

90.66 
1.73 



0.78 
o\83 



100.00 

13,980 
14,194 



52.5 
12.9 



ii 



3.45 

2.75 

87.90 

5.90 



100.00 

88.86 
2.04 
0.90 
1.95 
0.35 
5.90 



100.00 

13,950 
14,103 



42.5 
32.0 






t 
1.37 
3.59 

89.11 
5.93 



100.00 

87.70 
2.56 
1.03 
2.26 
0.56 
5.89 



100.00 



14,217 



34.4 
29.9 






t 

1.97 

4.35 

86.49 

7.19 



100.00 

85.66 
2.78 
0.77 
2.87 
0.64 
7.28 



100.00 



14,038 



30.9 
11.0 



So 

CO 



t 

2.08 

7.27 

74.32 

16.33 



100.00 

75.21 

2.81 
0.80 
4.08 
0.77 
16.33 



100.00 

12,472 
12,426 



26.7 
10.2 



a >> 



t 

1.50 

7.84 

81.07 

9.59 



100.00 

83.20 
3.29 
0.95 
2.45 
0.50 
9.61 



100.00 



14,003 



25. 

10.4 



* Authority not stated. f H. J. Williams. t U. S. Geological Survey. 

For smaller sizes a thinner bed has to be carried unless a strong 
draft is used. There is difficulty in keeping a thin bed free from air- 
holes. When possible the coal should be at least six inches deep on 
the grate. On account of the large percentage of ash in the smaller 






FUELS AND COMBUSTION 



19 



sizes, the fire requires frequent cleaning. Anthracites do not require 
" slicing" and should be disturbed only when cleaning is necessary. 
Nearly all anthracites, with some unimportant exceptions, come from 
three small fields in eastern Pennsylvania. On account of the limited 
supply and the great demand for domestic purposes, sizes over "pea 
coal" are prohibitive in price for steam power plant use. Table 1 
gives the composition and classification of a number of typical American 
anthracite coals, and Table 2, one of the standard divisions of mesh 
according to which L they are classed and marketed. Specific gravity, 
1.4 to 1.6. 

Burning No. 3 Buckwheat: Power, Dec. 27, 1910; Mar. 21, 1911. Burning 
Anthracite Culm of Poor Quality: Trans. A.S.M.E., 7-390. Anthracite Culm Bri- 
quets, Am. Inst. Min. Engrs., Bulletin, Sept., 1911. Calorific Value of Anthracite: 
Mines and Minerals, Sept., 1911. Preparation of Anthracite: Am. Inst. Min. 
Engrs., Bulletin, Oct., 1911. 

TABLE 2. 

STANDARD SIZES OF MESH: ANTHRACITE COAL. 
(Lehigh Coal and Navigation Company, Philadelphia, Pa.) 



Steamboat 

Broken 

Egg 

Stove 

Nut 

Pea 

Buckwheat No. .1 . 
Buckwheat No. 2. 
Buckwheat No. 3 



Over 


U 


Round. 


Through 


4* 


Round. 


Over 


3* 


Round. 


Through 


3i 


Round. 


Over 


2A 


Round. 


Through 


2fk 


Round. 


Over 


1+* 


Round. 


Through 


m 


Round. 


Over 


14 
T6 


Round or £§ square 


Through 


14 

T6 


Round or |f square 


Over 


9 


Round. 


Through 


fk 


Round. 


Over 


5 
T6 


Round. 


Through 


A 


Round. 


Over 


3 


Round. 


Through 


fk 


Round. 


Over 


A 





12. Semi-anthracites. — These coals kindle more readily and burn 
more rapidly than the anthracites. They require little attention, 
burn freely with a short flame and yield great heat with little clinker 
and ash. They are apt to split up on burning and waste somewhat in 
falling through the grate. They swell considerably, but do not cake. 
They have less density, hardness and metallic luster than anthracite, 
and can generally be distinguished by their tendency to soil the hands, 
while pure anthracite will not. Semi-anthracites are not of great im- 
portance in the steam power plant field on account of the limited sup- 
ply and high cost. They are found in a few small areas in the western 
part of the anthracite field. Specific gravity, 1.3 to 1.4. 



20 



STEAM POWER PLANT ENGINEERING 



13. Semi-bituminous. — These coals are similar in appearance to 
semi-anthracite, but they are somewhat softer and contain more vola- 
tile matter. They have a very high heating value, have a low moisture, 
ash and sulphur content, are readily burned without producing ob- 
jectionable smoke and rank among the best steaming coals in the 
world. The supply is limited and on account of high cost, except in 
the immediate vicinity of the mines, they are not generally used for 
power purposes. Table 3 gives the composition and classification of 
a number of typical American semi-anthracite and semi-bituminous 

coals. 

TABLE 3. 

COMPOSITION OF TYPICAL AMERICAN SEMI-ANTHRACITE AND SEMI-BITUMI- 

NOUS COALS. 



Proximate analysis: 

Water 

Volatile matter . . 

Fixed carbon 

Ash 



Ultimate analysis: 

Carbon 

Hydrogen 

Nitrogen 

Oxygen 

Sulphur 

Ash 



Calorific value: 

Calorimeter 

Dulong's formula 

Classification: 

Carbon-hydrogen ratio || 
Fuel ratio 



1.57 

9.40 

83.69 

5.34 



100.00 

85.46 
3.72 
1.12 
3.45 
0.91 
5.34 



100.00 



14,552 

23.0 

8.5 






2.36 

12.68 
72.88 
12.08 



100.00 

76.44 
3.82 
1.37 
4.30 
1.99 

12.08 



100.00 

13,259 
13,273 

20.7 
5.7 



r 



t 

4.07 
16.34 
68.47 
11.12 



100.00 

76.51 
4.27 
1.00 
6.59 
0.51 

11.12 



100.00 

13,509 
13,329 

19.6 
4.2 






X 

1.42 

20.72 
70.05 

7.81 



100.00 

81.95 
4.30 
1.29 

3.68 
0.97 

7.81 



100.00 

14,686 
14,363 

19.0 
3.4 



I* 



t 

1.53 
21.54 

71.88 
5.05 



100.00 

82.87 
4.76 
1.68 
4.99 
0.65 
5.05 



100.00 

14,807 
14,691 

17.8 
3.3 



0.44 
18.76 
73.15 

7.65 



100.00 

80.32 
4.88 
1.46 
4.69 
1.00 
7.65 



100.00 



14,432 

16.5 
3.9 



* Authority not stated. f U. S. Geological Survey. X W. Va. Geological Survey. 
§ H. J. Williams. || Based on air-dried sample. 

14. Bituminous. — These coals are the most widely distributed and 
the most extensively used fuel in steam power plant engineering. They 
contain a large and varying amount of volatile matter and burn freely 
with the production of considerable smoke unless carefully fired. Their 
physical properties vary widely and they are commonly classified as 

1. Dry, or free-burning bituminous. 

2. Bituminous caking. 

3. Long-flaming bituminous. 



FUELS AND COMBUSTION 



21 



•UMOJ 






'euquiBQ 



•III 






3 



3 



•T3M.OI 

^^pupp^ 






'3J99J03SJOJJ 



'PPkl 



•pnj 



O 00 00 05^00 



8 



3 



3 



8 



•qoiK 



•°N0 
Sinjpojj 



^A - M 



3 



8 



op t^ »C ■"* t>- CO 



3 



3 



O (NCO 
<N h 



8 ©© 2^ 



3 



0(M 

iCCO 

coco 

CiCS 



3 



o 3 



CO CO 

00 to 
iO CO 



8 



3 



ON 
tO <M 

00 Oi 



3 



O t-h 00 
-# O 



O CO -* 
O^ 

8 eo<N~ 



O OlN 

o cocq 

OOi 

8 ^"^f 



as 



i o 



-m o 



bfi a> 
O M 



88 

oi d 



TjH CO O 1 

CO CO CO ' 

ooo 



8 



I S3" 9 " 



N i— I OI 

(N 00 *# ■ 

OOO 



<M t-- O 



00 to o 

^ iO CO 



8 nn 3 



O CO 

cdci 



c3 fl 
■-2 +=«*-! O ?> 



c3 



2 ^ 



.5 ^u^L^ X 3 oq 



|l> 



O s3 






72 £ 



O ^ 5 on ^ 






o 



N 



° a 

€ S 

s-a 

9 V. 



n 



22 STEAM POWER PLANT ENGINEERING 

1. Dry bituminous coals are the best of the bituminous variety for 
steaming purposes. They are hard and dense, black in color, but 
somewhat brittle and splintery. They ignite readily, burn freely with 
a short clean bluish flame and without caking. Specific gravity, 
1.25 to 1.40. 

2. Bituminous caking coals swell up, become pasty and fuse together 
in burning. They contain less fixed carbon and more volatile matter 
than the free-burning grades. Caking coals are rich in hydrocarbon 
and are particularly adapted to gas making. The flame is of a yellow- 
ish color. Specific gravity, about 1.25. 

3. Long-flaming bituminous coals are similar in many respects to 
the caking coals but contain a larger percentage of volatile matter. 
They burn freely with a long yellowish flame. They may be either 
caking, non-caking or splintery. They are very valuable as a gas coal, 
and are little used for steaming purposes. Specific gravity, about 1.2. 

Table 4 gives the composition and classification of a number of 
typical American bituminous coals. 

For sizes of bituminous coal see paragraph 31. 

COAL FIELDS OF THE UNITED STATES. 

Alabama: Mines and Minerals, May, 1901. 

Alaska: Mining World, Aug. 28, 1909; Eng. and Min. Jour., Aug. 6, 1910. 

Arizona: Trans. Am. Inst. Min. Engrs., Feb. and May, 1902. 

California: Eng. and Min. Jour., July 4, 1896; Mining World, Feb. 17, 1906. 

Colorado: Mines [and Minerals, May, 1905, May, 1910; Min. Rept., Jan. 19, 1905. 

Idaho: Eng. and Min. Jour., March 5, 1910; Mines and Mining, Jan., 1903. 

Illinois: Min. Mag., March, 1905; Eng. and Min. Jour., Jan. 13, 1906. 

Iowa: Mines and Mining, Sept., 1910; Eng. and Min. Jour., May 10, 1902. 

Kentucky: Eng. and Min. Jour., Apr. 27, 1907; Jan. 18, 1908; June 27, 1908; 
Aug. 14, 1909. 

Maryland: U. S. Geol. Survey, Annual Report, 1902, part 3. 

Michigan: Eng. and Mining Jour., June 30, 1900; Min. World, Feb. 9, 1907. 

Mississippi: Eng. and Min. Jour., Jan. 16, 1909. 

Missouri: Am. Inst, of Min. Engrs., Jan., 1905. 

Montana: Min. Mag., March, 1905; Min. World, Nov. 24, 1906. 

Nebraska: Eng. and Min. Jour., Vol. 73, p. 481. 

Nevada: Eng. and Min. Jour., Dec. 31, 1910. 

New Mexico: Eng. and Min. Jour., May 21, 1910; June 20, 1908; March 7, 1908. 

North Carolina: Eng. and Min. Jour., June 11, 1910; Aug. 25, 1906. 

North Dakota: Eng. and Min. Jour., Jan. 15, 1910; April 10, 1909. 

Ohio: Min. Mag., Mar., 1905; U. S. Geol. Survey, part 3, 1902. 

Oklahoma (Ind. Terr.): Min. Rept., May 17, 1906; Mining World, Dec. 12, 
1908. 

Oregon: Eng. and Min. Jour., Aug. 17, 1907; Feb. 15, 1902. 

Pennsylvania: Eng. and Min. Jour., Aug. 24, 1901; Pro. Eng. S. W. Penn., Jan., 
1907. 

Rhode Island: Ry. Age Gazette, July 8, 1910. 



FUELS AND COMBUSTION 



23 



Tennessee: Eng. and Min. Jour., April 1, 1911; Mines and Mining, Sept., 1910. 
Texas: Mines and Minerals,. Oct., 1905. 
Utah: Mines and Mining, Sept., 1906; June, 1909. 

Virginia: Mines and Minerals, March, 1906; Eng. News, Oct. 20, 1904. 
Washington: Eng. and Min. Jour., Aug. 19, 1911; U. S. Geol. Survey, part 3, 
1902. 

West Virginia: Eng. and Min. Jour., May 12, 1904. 
Wisconsin: Trans. Am. Inst, of Mech. Engrs., Vol. 8, p. 478. 
Wyoming: Mining World, May 6, 1905; Coal Age, Apr. 13, 1912. 

GENERAL. 

Coal Mines of the United States: Peabody Atlas, A. Bement, Chicago, 111., Min. 
World, May 6, 1905; Eng. and Min. Jour., Jan. 8, 1910. 

Coal Resources of the Pacific: Eng. Mag., May, 1902. 

Rocky Mountain Coal Fields: Min. Rept., Jan. 5, 1905; Jour. Asso. Eng. Soc, 
Dec, 1902. 

Coal Fields, U. S. Northwest: Rev. of Rev., Feb., 1903. 

Coal Fields, U. S. Southwest: Eng. and Min. Jour., Oct. 17, 1903. 

Report of Coal Testing Plant: U. S. Geological Survey, Washington, D. C. (1906). 

Index of Mining Engineering Literature: W. R. Crane, John Wiley & Sons. 

15. Lignite, or brown coal, is a substance of more recent geological 
formation than coal and represents a stage in development intermediate 
between coal and peat. Its specific gravity is low, 1.2, and when 
freshly mined contains as high as 50 per cent of moisture. It is non- 
caking, and on exposure to air, slackens or crumbles. The lumps 
check and fall into small irregular pieces with a tendency to separate 
into extremely thin plates. It deteriorates greatly during storage or 
long transportation. Lignite, as mined, is a low-grade fuel with a 
calorific value of about one-half that of good coal. When properly 
prepared and compressed into briquettes lignite becomes an excellent 
fuel, resists weathering satisfactorily, permits handling and trans- 
portation without excessive deterioration and is practically smokeless. 
The superiority of briquettes over raw lignite is shown by the fol- 
lowing table: 

IMPROVEMENT OF HEAT VALUE BY BRIQUETTING.* 





Moisture 


Heat Value per Pound. 


Source 


In Raw 
Lignite. 


In Bri- 
quettes. 


Removed. 


Raw 
Lignite. 


Briquettes. 


Increase. 


Texas 


Per Cent. 
33.0 
40.0 
42.0 
40.0 


Per Cent. 

9.0 

12.0 

10.0 

10.0 


Per Cent. 
24.0 
28.0 

32.0 
30.0 


B.t.u. 

6840 
6241 
6079 
6080 


B.t.u. 
9336 

9354 
9355 
9264 


Per Cent. 

36.5 


North Dakota 

North Dakota 

California 


50.0 
54.0 
52.4 







Bulletin No. 14, U. S. Bureau of Mines, p. 48. 



24 



STEAM POWER PLANT ENGINEERING 



The most extensive lignite deposits are situated long distances from 
fields of high-grade coal, and although the use of lignite is at present 
limited to these regions it is fast becoming a general competitor of coal. 

North Dakota Lignite as a Fuel for Power Plant Boilers: Bui. No. 2, 1910, U. S. 
Bureau of Mines. Briquetting Tests of Lignite: Bui. No. 14, 1911, U. S. Bureau of 
Mines. General data pertaining to lignite fuels, Engr. U. S., Jan., 1910. 

TABLE 5. 

COMPOSITION AND CLASSIFICATION OF TYPICAL AMERICAN LIGNITES.* 

(Run of Mine.) 



Proximate analysis: 

Water 

Volatile matter 

Fixed carbon 

Ash 

Ultimate analysis: 

Hydrogen 

Carbon 

Nitrogen 

Oxygen 

Sulphur 

Ash 

Calorific value: 

Calorimeter 

Dulong's formula 

Classification: 

Carbon-hydrogen ratio f 

Fuel ratio 






11.05 
35.90 
42.08 
10.97 



100.00 

5.37 
59.08 

1.33 
21.52 

1.73 
10.97 



100.00 

10,539 
10,355 

11.50 
1.17 



§■8 j* 



12.29 

34.58 

46.14 

6.99 



100.00 

5.82 
63.31 

1.03 
22.22 

0.63 

6.99 



100.00 

11,252 
11,153 

11.20 
1.09 



Hpq 



33.71 
29.25 
29.76 

7.28 



100.00 

6.79 
45.52 

0.79 
42.09 

0.53 

7.28 



100.00 

7348 
7177 

10.90 
1.02 



6JS 

offl 



18.68 

34.88 

40.45 

5.99 



100.00 

6.07 

57.46 

1.15 

28.78 
0.55 
5.99 



100.00 

10,143 
9,948 



1.16 



m 



36.78 

28.16 

29.97 

5.09 



100.00 

6.93 
41.87 

0.69 
44.94 

0.48 

5.09 



100.00 

7002 
6944 

9.60 
1.06 



gj 



22.63 

35.68 

37.19 

4.50 



100.00 

6.39 
54.91 

1.02 
32.59 

0.59 

4.50 



100.00 

9734 

9478 

9.40 
1.05 



* Compiled from Government Report, U. S. Geological Survey. 
t Based on air-dried analysis. 



16. Peat, or Turf, is formed by the slow carbonization under water 
of a variety of accumulated vegetable materials. It is unsuitable for 
fuel until dried. Peat, as ordinarily cut and dried, is too bulky for 
commercial competition with coal, and is used only where coal is pro- 
hibitive in price. When properly prepared and compressed into bri- 
quettes peat is an excellent fuel. In Russia, Germany, and Holland 
peat briquettes have passed the experimental stage and several millions 
of pounds are manufactured annually. Peat is used but little in this 
country at present, though the deposits are extensive and widely dis- 
tributed, but its possibilities are beginning to attract the attention of 



FUELS AND COMBUSTION 



25 



engineers. The proportion in which the various primary constituents 
exist in dried peat is approximately as follows : 

Per Cent. 

Fixed carbon 35 

Volatile matter 60 

Ash 5 

Peat: Prac. Engr. U. S., Jan., 1910, p. 21; Bui. No. 16, U. S. Bureau of Mines, 
1911; Power, Sept. 6, 1910; Eng. and Min. Jour., Nov. 22, 1902; Feb. 7, 1903, 
Jour. Am. Peat Soc, July, 1911; Elec. Rev., Mar. 22, 1912; Min. and Eng. Wld., 
Nov. 28, 1911. 

17. Wood, Straw, Sawdust, Bagasse, Tanbark. — In certain localities 
cord wood is still used as a fuel, but the steadily increasing values 
of even the poorest qualities are rapidly prohibiting its use for steam- 

TABLE 6. 

PHYSICAL AND CHEMICAL PROPERTIES OF WOODS, STRAW AND TANBARK. 
(Prac. Engr. U. S., Jan., 1910.) 





o 
3 . 

3 K 

O -o 

n 5 

0> p 

bo o 


•6 

6 

b0 

g 


Equivalent Weight 
of Coal. 13,500 
B.T.U. 


6 

J* 

8 


a 

be >-, 
? & 

>> 


c 

6 

boPn 

X 
O 


bo S 
g ^ 


6 
.& 

< 


Calorific value, 
B.T.U. per 
Pound. 


'1 
< 


Ash 


46 
43 
45 
42 
41 
35 
25 
53 
49 
59 
52 
45 
25 
36 
36 
25 
35 
25 


3520 
3250 
2880 
3140 
2350 
2350 
1220 
4500 
3310 
3850 
3850 
3310 
1920 
2130 
2130 
1920 
3310 
1920 


1420 

1300 

1190 

1260 

940 

940 

580 

1800 

1340 

1560 

1540 

1340 

970 

1050 

1050 

970 

1340 

970 












5450 
5400 
5580 
5420 
5400 
54Q0 
6410 
5400 
5460 


Hutton 


Beech ... 
Birch 




49.36 
50.20 


6.01 
6.20 


42.69 
41.62 


0.91 
1.15 


1.06 
0.81 


Sharpless 
Hutton 


Cherry . . . 




u 












Sharpless 


Elm 








































Sharpless 


Maple, Hard 






















5460 
5400 
5460 
6830 
6660 


it 


" White. 
" Red 


49.64 


5.92 


41.16 


1.29 


1.97 


Rankine 
Hutton 


Pine, White 
" Yellow 












tt 












tt 


Poplar 

Spruce 

Walnut 


49.37 


6.21 


41.60 


0.96 


i.86 


6660 
6830 
5460 
6830 


it 
tt 












tt 


Willow .... 


49.96 
49.70 


5.96 
6.06 


39.56 


0.96 
1.05 


3.37 
1.80 


Rankine 


Average. 




41.30 














Straw : 
Wheat . . . 
Barley . . . 


* 

00 

s 




Water 
16.00 
15.50 


35.86 
36.27 


5.01 

5.07 
5.04 


37.68 
38.26 


0.45 
0.40 
0.42 


5.00 
4.50 
4.75 


5155 


Clark 


Average 


15.75 


36.06 


37.97 




Tanbark 

Drv 








51.80 


6.04 


40.74 




1.42 


6100 


Myers 













Compressed. 



26 



STEAM POWER PLANT ENGINEERING 



generating purposes. Sawdust, shavings, tanbark and other waste 
products of wood are burned under boilers in situations where such 
disposition nets the best financial returns. Recent progress, however, 
in industrial chemistry shows that ethyl and wood alcohols and other 
valuable by-products can be cheaply made from sawdust, shavings, 
slashings and similar waste material, and it is not unlikely that their 
use for steaming purposes will be unheard of in a comparatively few 
years. Table 6 gives the physical and chemical characteristics of a 
number of woods. 

Wood as Fuel: Prac. Engr. U. S., Jan., 1910, p. 805; Power & Engr., June 30, 
1908, p. 1015; Power, Dec, 1908, p. 772. 

Burning Sawdust: Prac. Engr. U. S., Jan., 1910, p. 48; Power & Engr., April 7, 
1908, p. 536; Oct. 13, 1908, p. 613; Jour, of Elec, Oct., 1905. 

TABLE 7. 

HEAT VALUES OF BAGASSE AND VARIATION WITH DEGREE OF EXTRACTION. 



Si 

12 




Fiber. 


Sugar. 


Molasses. 


£6 

W . 
o o 


Heat required to 

evaporate the Water 

present. B.T.U. 


6 

c3 P3 


Lbs. Bagasse required 
to equal lib. Coal of 
14,000 B.T.U. Cal- 
orific Power. 


Coal Equivalent per 

Ton of Cane. 

Pounds. 


0) 

o 


H 


c 

il 

fed 


It 


d 

o & 

s-, c3 
& CQ 


1 • 
Spa 


S 23 




£ fa 

g 
0) 


"on 


0.00 

28.33 


100.00 
66.67 


8325 
5552 










8325 
5900 


339 


8325 
5561 


1.68 
2.52 


119 
119 


2465° 


85 


3.33 


240 


1.67 


116 


2236 


80 


42.50 


50.00 


4160 


5.00 


361 


2.50 


174 


4697 


509 


4188 


3.34 


120 


2023 


75 


51.00 


40.00 


3330 


6.00 


433 


3.00 


209 


3972 


611 


3361 


4.17 


120 


1862 


70 


56.67 


33.33 


2775 


6.67 


482 


3.33 


232 


3489 


679 


2810 


4.98 


120 


1732 


65 


60.71 


28.57 


2378 


7.15 


516 


3.57 


248 


3142 


727 


2415 


5.80 


121 


1612 


60 


63.75 


25.00 


2081 


7.50 


541 


3.75 


261 


2883 


764 


2119 


6.61 


121 


1513 


55 


66.12 


22.22 


1850 


7.78 


562 


3.88 


270 


2682 


792 


1890 


7.40 


121 


1427 


50 


68.00 


20.00 


1665 


8.00 


578 


4.00 


278 


2521 


815 


1706 


8.21 


122 


1350 


45 


69.55 


18.18 


1513 


8.18 


591 


4.09 


284 


2388 


833 


1555 


9.00 


122 


1284 


40 


70.83 


16.67 


1388 


8.33 


601 


4.17 


290 


2279 


849 


1430 


9.79 


123 


1222 


25 


73.67 


13.33 


1110 


8.67 


626 


4.33 


301 


2037 


883 


1154 


12.13 


124 


1077 


15 


75.00 


11.77 


980 


8.82 


637 


4.41 


307 


1924 


899 


1025 


13.66 


124 


1002 





76.50 


10.00 


832 


9.00 


650 


4.50 


313 


1795 


916 


879 


15.93 


126 


906 



Bagasse, or megass, is refuse sugar cane and is used as a fuel on the 
sugar plantations. Its heat value depends upon the proportions of 
fiber, molasses, sugar and water left after the extraction. The heat 
furnished by the different constituents is about as follows: Fiber, 
8325 B.t.u. per pound; sugar, 7223 B.t.u. per pound; and molasses, 
6956 B.t.u. per pound. Table 7 gives the heat value of bagasse and 
variation with the degree of extraction. A typical furnace for burning 
bagasse is shown in Fig. 5. 

Bagasse as Fuel: Prac. Engr. U. S., Jan., 1910; Engng., Feb. 18, 1910. 

Bagasse Drying: E. W. Kerr, Louisiana Bui. No. 128, June, 1911. 



FUELS AND COMBUSTION 





28 



STEAM POWER PLANT ENGINEERING 




FUELS AND COMBUSTION 29 

Tanbark is usually quite moist; the amount of moisture varies with 
the Jeaching process used and averages around 65 per cent. In this 
condition it has a heat value of about 4300 B.t.u. per pound. If per- 
fectly dry its heating power is approximately 6100 B.t.u. per pound. 
As in the case of all moist fuels, tanbark must be surrounded by heated 
surfaces of sufficient extent to insure drying out the fresh fuel as thrown 
on the .fire. A successful furnace for burning tanbark is shown in 
Fig. 6. 

Tanbark as a Boiler Fuel: Jour. A.S.M.E., Feb., 1910, p. 181; Jour. A.S.M.E., 
Oct., 1909, p. 951; Prac. Engr. U. S., Jan., 1910. 

18. Combustion. — By combustion is meant the chemical union of 
the combustible material of a fuel and the oxygen of the air. Theo- 
retically the process is a simple one, as it is only necessary to bring 
each particle of fuel previously heated to the kindling temperature in 
contact with the correct amount of oxygen and the combustion will be 
complete, the fuel oxidizing to the highest possible degree. In practice, 
however, the size and character of fuel, type of furnace, draft, impuri- 
ties in the fuel, and the mechanical difficulties affect combustion to 
such an extent as to render oxidation more or less incomplete. 

When heat is applied to coal, combustion takes place in three separate 
and distinct stages: 

1. Absorption of heat. A fresh charge of fuel when thrown on a 
fire must first be brought to the kindling point in order that chemical 
action may take place. The temperatures necessary to cause this 
union of oxygen and fuel are approximately as follows: 

Degrees F. Degrees F. 

Lignite dust 300 Cokes 800 

Sulphur 470 Anthracite lump 750 

Dried peat 435 Carbon monoxide 1211 

Anthracite dust 570 Hydrogen '. 1100 

Lump coal 600 

(Stromeyer, Marine Boiler Management and Construction, p. 93.) 

2. Vaporization of the hydrocarbon portion of the fuel and its com- 
bustion, the hydrocarbons consisting principally of defiant gas, C2H4, 
marsh gas, CH 4 , tar, pitch, naphtha and the like. As these gases are 
driven off they become mixed with the entering air, and the carbon 
and hydrogen unite with the oxygen, forming carbon dioxide, C0 2 , and 
water vapor, H 2 0, respectively, and give up heat in doing so. If 
volatile sulphur is present it unites with the oxygen, forming sulphur 
dioxide, S0 2 , and also gives up heat, but its presence is objectionable, 
as the S0 2 , particularly, in the presence of moisture, attacks the metal 
of the furnace and boiler and causes rapid corrosion. If insufficient 



30 STEAM POWER PLANT ENGINEERING 

oxygen is present for complete oxidation, the carbon may burn to carbon 
monoxide, CO, and only a small portion of the available heat be liberated. 

3. Combustion of the solid or carbonaceous portion of the fuel. 
After the hydrocarbons have been driven off and oxidized the remain- 
ing solid matter is composed chiefly of carbon and ash. The carbon 
unites with the oxygen, forming carbon dioxide, carbon monoxide, or 
both, depending upon the completeness of combustion. The ash, of 
course, remains unconsumed. 

In commercial practice the requirements for perfect combustion are 
a surplus of air, a thorough mixture of the fuel particles with the air, 
and a high temperature. The surplus of air above theoretical require- 
ments should be kept to a minimum, but even in the most scientifically 
designed furnace some excess is essential on account of the difficulty 
of properly mixing the gases, since the currents of combustible gases 
and air are apt to be more or less stratified. The products of com- 
bustion must be maintained at the kindling temperature until oxidation 
is complete, otherwise the carbon will be wasted as carbon monoxide 
or as smoke. The final products of combustion as exhausted by the 
chimney should consist only of carbon dioxide, water vapor, oxygen, 
nitrogen, and the oxides of impurities in the fuel. 

Prof. Wm. A. Bone of the University, Leeds, England, has recently 
advocated "nameless incandescent surface combustion" as a means of 
greatly increasing the general efficiency of industrial furnaces wherever 
it can be conveniently applied. (Eng. News, Jan. 18, 1912, p. 96; 
Engineering, April 14, 1911.) 

"The distinguishing and essential feature of the new process is 
that a homogeneous explosive mixture of gas and air, in the proper 
proportions for complete combustion, is caused to burn without flame 
in contact with a granular incandescent solid, whereby a large portion 
of the potential energy of the gas is immediately converted into radiant 
form." Prof. Bone claims that he has been able to transmit 95 per 
cent of the available energy of gaseous fuel to the water in specially 
constructed fire-tube boilers. For a description of the apparatus used 
see paragraph 55. 

When the combustible elements unite with oxygen they do so in 

definite proportions called the molecular weights, which are always the 

same, and the union produces a fixed quantity of heat. Thus, in the 

combustion of carbon to CO, 24 pounds of carbon unite with 32 pounds 

of oxygen, forming 56 pounds of CO; hence one pound of carbon will 

form 

C 2 + 2 24+32 OQ . , fnr . 

-~ — - = — ^ — = 2.34 pounds of CO. 
C2 2i<± 



FUELS AND COMBUSTION 31 

The heat of combustion will be 4380 B.t.u. per pound of carbon thus 
consumed. 

In the burning to C0 2 one pound of carbon will form 

° 2 "t 2 ° 2 = 24 + <? A X 32 = 3| pounds of C0 2 and liberates 14,540 B.t.u. 

C 2 -"4 

Similarly, in burning to H 2 one pound of hydrogen will form 
2H 2 + 2 2X2 + 32 . , ._ n 

2H 2 = 2X2 = 9 P ° Unds ° f H2 °' 

(The exact figures, based upon the relative molecular weights, as 
adopted by the International Committee on Atomic Weights, are 
2 X 2.016 + 32 



2 X 2.016 



= 8.94 pounds. 



For all practical engineering purposes the use of the exact values of 
the molecular weights is an unnecessary refinement and the decimal 
factors may well be omitted. In the ensuing calculations only the ap- 
proximate values will be considered.) If the products of combustion 
are condensed and their temperature lowered to the initial temper- 
ature of the constituent gases the heat liberated will be 61,950 B.t.u. 
This is known as the higher heating value. If the products of com- 
bustion are not condensed, which is the usual case in practice, the 
latent heat of vaporization of the water vapor is not available. The 
difference between the higher heating value and the unavailable heat 
is called the net, or lower heating value. The unavailable portion of the 
heat depends upon the temperature at which the products of com- 
bustion are discharged. This varies with practically every installa- 
tion. Thus, one pound of water vapor escaping uncondensed in the 
products of combustion at temperature h degrees F., will carry away 
approximately (1090.7 + 0.455 t x — t) B.t.u. above initial tempera- 
ture t degrees F. of the constituent gases. (Carpenter & Diederichs, 
Exp. Engng., 1911, p. 467.) Since one pound of hydrogen burns to ap- 
proximately 9 pounds of water vapor, the lower heating value h f will 
be 

h! = 61,950 - 9 (1090.7 + 0.455 h - t) B.t.u., (0) 

Many attempts have been made to adopt a standard lower heating- 
value, but the results have been far from harmonious. In a letter 
dated Jan. 12, 1912, and addressed to the author, the U. S. Bureau of 
Standards recommends "that the quantity to be subtracted from the gross 
value to give the net value be taken as the latent heat of vaporization at 
degrees Centigrade, of the water r formed during combustion, and of that 
contained in the fuel." 



32 



STEAM POWER PLANT ENGINEERING 



This would give the net or lower heat value of hydrogen as 

61,950 - 9 (1073.4) = 52,290 B.t.u. 

For ti = t = 0° C. = 32° F., formula (0) gives the same result. 

In the ordinary furnace the oxygen is obtained from the atmosphere 
which, neglecting moisture and a few minor elements, contains the 
following, mechanically mixed : 





By Volume. 


By Weight. 


Nitrogen 

Oxygen 


79.04 
20.96 


76.80 
23.20 





For most engineering purposes this relationship may be expressed ; 





By Volume. 


By Weight, 


Nitrogen 


79 
21 


77 
23 


Oxygen 





Hence, in the combustion of one pound of pure carbon the products 
of combustion contain not only 3| pounds of C0 2 , but \\ X 2f = 8.92 
pounds of nitrogen, giving a total of 3| + 8.92 = 12.58 pounds. The 
nitrogen performs no useful office in combustion and is supposed to 
pass through the furnace without change. It dilutes the products of 
combustion and reduces the temperature. 

Table 8 gives the physical and chemical properties of the substances 
most commonly met with in connection with combustion. 

19. Calorific Value of Coal. — The heat liberated by the com- 
bustion of unit weight of fuel is called the calorific value of the fuel. 
The most rational way of determining the heat of combustion is to 
burn a weighed sample of coal in an atmosphere of oxygen in a suit- 
able calorimeter. An alternative method is to calculate the heat of 
combustion from the chemical analysis. An analysis which determines 
the per cent of fixed carbon, volatile matter, moisture, and ash, is called 
the proximate analysis, while one wmich reduces the fuel to its elemen- 
tary constituents of carbon, hydrogen, nitrogen, sulphur, moisture, and 
ash is called the ultimate analysis. The proximate analysis is com- 
paratively easy to make and gives the general characteristics of a fuel. 
It is made by subjecting the sample to a moderate temperature to 
expel the moisture, then to a higher temperature until the volatile 
matter is driven off, and finally to a very high temperature which 
drives off all carbon as carbon dioxide and leaves the ash as a residue. 



FUELS AND COMBUSTION 



33 



By weighing the residue at the end of each operation the various per- 
centages may be computed. For method of making proximate analysis, 
see " Report of Committee on Coal Testing," Journal of the American 
Chemical Society, Vol. 21, p. 1116. For method of making both prox- 
imate and ultimate analyses, see " Report of Coal Testing Plant," 
U. S. Geological Survey, No. 48, Part II, 1906, and "Method of Analyz- 
ing Coal and Coke," U. S. Bureau of Mines, Technical Paper, No. 8, 

1912. 

TABLE 8. 

DATA RELATIVE TO ELEMENTS MOST COMMONLY MET WITH IN CONNECTION WITH 

COMBUSTION OF FUEL. 



Substance. 


Molecular 
Formula. 


Relative 

Molecular 

Weight, 

Oxygen 

=32. 


Chemical Reactions. 


Weight per 
Pound of Sub- 
stance in First 
Column. 




Oxygen. 


Air. 


Acetylene 

Air. . . ... 


C2H2 


26.02 


2 C 2 H 2 +5 2 = 4 C0 2 +2 H 2 


3.08 


13.35 


Ash 












Carbon 


c 2 
c 2 

co 2 

CO 
H 2 
CH 4 

N 2 
C 2 H4 

o 2 

S 2 
H 2 


24.0 
24.0 
44.0 
28.0 
2.016 
16.03 
28.02 
28.03 
32.0 
32.07 
18.02 


2C+0 2 = 2CO 
2C+2 2 = 2C0 2 


1.33 

2.66 


5.78 


Carbon 

Carbon dioxide .... 


11.58 


Carbon monoxide.. . 

Hydrogen 

Marsh gas 

Nitrogen 


2CO+0 2 = 2C0 2 
2H 2 +0 2 = 2H 2 
CH 4 +2 2 = C0 2 +2H 2 


0.57 

8.0 

4.0 


2.47 
34.8 
17.4 


Olefiant gas 

Oxygen 

Sulphur 

Water vapor 


C 2 H 4 +3 2 = 2 CO2+2H2O 


3.43 


14.9 


S 2 +2 2 = 2S0 2 


1.0 


4.32 







Substance. 


Mean 

Specific 

Heat. 


Density and Specific Volume 
at 32° F., and 14.7 Lbs. per 
Sq. In.* 


Heat of Combustion 
(Higher Heat Value) 
B.t.u.f 


Weight per 
Cu. Foot. 


Cu. Feet 
per Lb. 


Per Pound. 


Per Cu. Foot 

at 32° F., and 

14.7 Lbs. 


Acetylene 


a 

a 
m 


0.0725 

0.0807 


13.79 
12.39 


21,430 


1582 


Air 




Ash 






Carbon 


145 (solid) 

145 (solid) 

0*1227 

0.0781 

0.0056 

0.0447 

0.0783 

0.0781 

0.0892 

125 (solid) 




4,380 
14,540 




Carbon 






Carbon dioxide .... 


8.15 
12.80 
177.9 
22.37 
12.77 
12.80 
11.21 




Carbon monoxide . . 
Hydrogen 


4,380 
61,950 
23,840 


342 
345 


Marsh gas 


1067 


Nitrogen 




Olefiant gas 

Oxygen 


21,450 


1685 


Sulphur 


4,020 




Water vapor 

















* Smithsonian tables. t Carpenter and Diederichs, Exp. Eng., 1911. 



34 STEAM POWER PLANT ENGINEERING 

The formula most commonly used in calculating the heating value 
of a fuel is based on the ultimate analysis and is known as Dulong's 
formula. This is based on the assumption that the oxygen in the fuel 
and enough hydrogen to unite with it may be considered inert and the 
remainder of the hydrogen and all the carbon and sulphur may be 
treated as free elements, thus: 

h d = 14,540 C + 61,950 (h - §) + 4020 S,* (1) 

in which 

hd = heating value in B.t.u. per pound of fuel. 

C, H, 0, and S refer to the proportions of carbon, hydrogen, oxygen, 
and sulphur, respectively, in the fuel. 

With fuels low in volatile matter values calculated by means of 
Dulong's formula agree closely with calorimetric determinations, but 
may be in considerable error for fuels having more than 20 per cent 
of volatile matter. 

For a comparison between the actual heat values and those cal- 
culated by means of Dulong's formula for various coals, see Tables 

1 to 5. 

The following modification of Dulong's formula, as developed by 
Mahler, gives results for Pennsylvania and Ohio coals agreeing within 

2 per cent of calorimeter determinations. (Trans. A.I.M.E., Feb., 
1897.) 

h m = 200 C + 675 H - 5400, (2) 

in which C and H are in per cent. 

The heating value of certain classes of coals may be estimated from 
the proximate analysis. Thus, for Illinois coals with ash content 
under 10 per cent, R. W. Hunt & Co. deduced the formula 

h h = 14,544 C + 16,515 V - 10,000 A, (3) 

in which 

h h = Bot.u. per pound of coal. 

C, V, and A = the proportional content of fixed carbon, volatile 
matter, and ash. 

When ash lies between 10 and 15 per cent, the formula will be more 
accurate if written 

h h = 14,544 C + 16,515 V + 354 A - 1635. (4) 

* Dulong's formula is usually stated: 

(a) Heating value per pound = 14,600 C + 62,000 (b. - ^) + 4000 S. 
The U. S. Geological Survey uses: 

(6) Heating value per pound = 14,544 C + 62,028 I H - 2\ + 4050 S. 



FUELS AND COMBUSTION 



35 



Kent ("Steam Boiler Economy," First Edition, p. 47) deduced from 
Mahler's tests of European coals the following relationship between 
the approximate heating value and the percentage of fixed carbon in 
the combustible: 



Percentage of Fixed 

Carbon in Coal, 

Dry and Free 

from Ash. 


Heating Value 
B.t.u. per Pound 
of Combustible. 


Percentage of Fixed 

Carbon in Coal, 

Dry and Free 

from Ash. 


Heating Value 
B.t.u. per Pound 
of Combustible. 


100 
97 
94 
90 

87 
80 

72 


14,600 
14,940 
15,210 
15,480 
15,660 
15,840 
15,660 


68 
63 
60 
57 
55 
53 
51 


15,480 
15,120 
14,580 
14,040 
13,320 
12,600 
12,240 



Goutal (Comptes Rendus de l'Academie des Sciences, Vol. 135, 
p. 477) gives carbon a fixed value and considers the heat value of the 
volatile matter a function of its percentage referred to combustible : 

hg = 14,760 C + aV, (5) 

in which 

h g = B.t.u. per pound of coal. 

C = proportional content of fixed carbon in the coal. 
V = proportional content of volatile matter in the coal. 
a = coefficient as per following table: 



V 




v 




v + c 




v + c 




05 


26,100 


.26 


18,360 


10 


23,400 


28 


17,980 


12 


22,350 


30 


17,640 


14 


21,450 


32 


17,300 


16 


20,750 


34 


17,000 


18 


20,220 


36 


16,680 


20 


19,620 


37 


16,180 


22 


19,220 


• 38 


15,300 


24 


18,790 


40 


14,400 



Kent's and Goutal's methods give results which are accurate enough 
for ordinary work when applied to eastern coals within their range, 
but they apply with less accuracy to the coals of the middle west and 
are quite unreliable for coals in the far west and north. 

Calorimetric determinations are necessary in all cases where accuracy is 
required. 



36 STEAM POWER PLANT ENGINEERING 

Tables 1 to 4 give the proximate and ultimate analyses and the 
calorimetric and calculated heat values for a number of American coals. 

For a means of determining the hydrogen, nitrogen, total carbon and 
oxygen content of coal from the proximate analysis, see "Ultimate 
Analysis of Coal," by Prof. L. S. Marks, Power, Dec, 1908, p. 928. 

Fuel Calorimeters: See paragraph 458. 

Calorific Value of Fuels: Engr., London, Feb. 17, 1911; U. S. Geological Survey, 
Bulletin Nos. 261, 290, 323, 325, 332; Jour. Franklin Inst., P. Mahler, Jan., 1905; 
Prac. Engr., U. S., Jan., 1910. 

Recent Progress in Calorimetry: Met. and Chem. Engrg., Sept., 1911; Com- 
parison of Calorimeters, Jour. Soc. Chem. Ind., 22-1230, 23,704. 

20. Air Required for Combustion. — One pound of carbon in burn- 
ing to C0 2 requires 2.66 pounds of oxygen or 2.66 -f- 0.23 = 11.58 
pounds of dry air. It may be shown in a similar manner that one 
pound of hydrogen requires 34.8 pounds of dry air. Since the com- 
bustible portion of all commercial fuels consists chiefly of carbon and 
hydrogen the theoretical air requirements may be approximated from 
the ultimate analysis as follows: 

A 1 = 11.58 C + 34.8 (h-^Y (6) 

in which 

Ai = weight of dry air required per pound of fuel, pounds. 
C, H and O = proportional part of dry weight of carbon, hydrogen 
and oxygen in the fuel. > 

'q- = proportional part of the hydrogen supplied with oxygen 

o 

from the fuel itself.* 
Equation (6) is commonly written: 

A=35 (C + H _0). ( 7 ) 

Example: Required the theoretical weight of dry air supplied per 
pound of coal as fired with analysis as follows : 

Per Cent. Per Cent. 

Carbon 65 Ash and Sulphur 13 

Hydrogen 5 Water 8 

Oxygen 8 Total 100 

Nitrogen 1 

* This term ( H — ^ ) does not contain a proper correction for the hydrogen 

contained in the moisture, for not all of the oxygen in coal is combined with hydro- 
gen. Part of the oxygen is probably combined with nitrogen in organic nitrates 
and part may be present in carbonates in mineral matter caught in the coal. The 
error of this assumption, however, lies within the accuracy of the average boiler 
observations. 




FUELS AND COMBUSTION 37 

Substituting the value of C, H, and O in equation (6) 

A x = 11.58 X 0.65 + 34.8^0.05 - ^~) = 8.92 pounds, 

the theoretical weight of dry air necessary to burn one pound of coal 
as fired. 

Since the coal contains 8 per cent of moisture the weight of dry air 
required per pound of dry coal is 

8 - 92 OAO A 

t-^ = 9.69 pounds. 

The water and ash only are treated as incombustible, therefore the 
air required per pound of combustible is 

^q = 11.29 pounds. 

Example: Required the character and amount of the products of 
combustion if one pound of coal, as per following analysis, is completely 
burned with the theoretical air requirements. 

Per Cent. Per Cent. 

Carbon 65 Ash 12 

Hydrogen 5 Water 8 

Oxygen 8 Sulphur 1 

Nitrogen 1 Total 100 

The products of combustion will consist of C0 2 , N 2 , H 2 0, ash, and 

possible S0 2 or S0 3 , thus: 

44 
The carbon will produce 0.65 X j^ = 2.38 lbs. of C0 2 



The air for the carbon will liberate . 0.65 X Vs X 7^7 = 5.80 lbs. of N 2 



32 77 
12^23 

The available hydrogen will produce . f 0.05 ^- j9 = 0.36 lbs. of H 2 

The air for the hydrogen will liberate 



(0.05 -*», 

the 



X § = L071bs - ofN 2 



The oxygen and inert portion of 
hydrogen will appear as vapor .... 0.08 -] — 75—-= 0.09 lbs. of H2O 

o 

The nitrogen in the fuel is considered inert * . . . . 0.01 lbs. of N 2 

The moisture will appear as vapor 0.08 lbs. of H 2 

The sulphur f is usually treated as ash . . 0.12 + 0.01 = 0.13 lbs. of ash 

Total products of combustion = 9.92 lbs. 

* This is not true since a large percentage of the nitrogen content of the fuel 
appears in the flue gas in combination with other elements, but the amount is so 
small compared with that supplied in the air that no appreciable error arises from 
the assumption that it remains inert and passes through the furnace without change. 

t The sulphur content is ordinarily so small that no attempt is made to separate 
the volatile and non-volatile constituents and the whole is treated as ash. If the 



38 STEAM POWER PLANT ENGINEERING 

The distribution of the elementary gases and compounds is as follows: 





C 2 


H 2 


2 


N 2 


Air. 


C0 2 


H 2 o 


Ash. 


C to CO* 


0.65 


0^04 
0.01 


1.73 
0.32 
0.08 


5.80 
1.07 


7.53 
1.39 


2.38 






Available H to H 2 


0.36 
0.09 




and inert H to H2O 






Nitrogen in fuel 




0.01 








Moisture in fuel 












0.08 




Ash 














0.13 




















Total 


0.65 


0.05 


2.13 


6.88 


8.92 


2.38 


0.53 


0.13 







The weight of gaseous products is 9.92 — 0.13 = 9.79 pounds, or from 

the table, 

0.65 + 0.05 + 2.13 + 6.88 + 0.08 = 9.79 pounds. 

The weight of dry gases is 

9.79 - (0.36 + 0.09 + 0.08) = 9.26 pounds. 

The weight of dry air supplied is, from the table, 8.92, which checks 
with the results as calculated from equation (6) . 

In practice the amount of air supplied is measured directly in situ- 
ations where such measurements can be readily made, or, as is usually 
the case, it is calculated from the flue-gas analysis. 

Air excess is essential in the commercial combustion of solid fuels, 
and the gaseous products of combustion will contain 2 and possible 
CO in addition to C0 2 , N 2 , H 2 0, and S0 2 , as obtained from perfect com- 
bustion with theoretical air supply. 

Example: Required the amount of dry air supplied per pound of 
coal, as per preceding analysis, if the dry flue gas resulting from the 
combustion is composed of 

C0 2 , 14 per cent by volume. 

CO, 0.5 per cent by volume. 

2 , 6.0 per cent by volume. 

N 2 , 79.5 per cent by volume. 

Temperature of sample, 62 degrees F., barometer 30 inches. 

The weights may be determined from the density given in Table 8. 





Volume X Density 


= ' Weight. 


C0 2 

2 


14 

6 

0.5 


0.1159 
0.0843 
0.0737 


1.6226 
0.5058 
0.0368 


CO 



volatile portion is to be considered the influence of the S0 2 or S0 3 in the flue gas 
should be included in the heat balance. Some engineers treat one-half the sulphur 
as volatile and the balance as ash. 



FUELS AND COMBUSTION 39 

These weights may be subdivided into those of their constituents, 
thus the C0 2 contains T 3 T of carbon and T 8 T of oxygen, and the CO, f of 
carbon and f of oxygen. 

T 8 T X 1.6226 = 1.1800 T 3 T X 1.6226 = 0.4426 

t X 0.0368 = 0.0210 f X 0.0368 = 0.0158 

Free oxygen = 0.5058 



Pounds of oxygen 1.7068 Pounds of carbon 0.4584 

1 7068 
Weight of oxygen per pound of carbon, ' = 3.72. 

U.4t/o4 

Since 23 per cent of air by weight is oxygen, weight of dry air per 

3 72 
pound of carbon = ^r- = 16.2. 

Since the coal used contains 65 per cent carbon, the weight of dry 
air supplied for the carbon is 

16.2 X 0.65 = 10.52 pounds. 
The hydrogen in the fuel required, 

34.8^0.05 - ^) = 1.39 pounds. 

The total weight of dry air supplied per pound of coal as fired is 
10.52 + 1.39 = 11.91 pounds. 

The total weight of the dry products of combustion per pound of 
coal as fired is 

10.52 + j^r X 1.39 + 0.65 + 0.01 = 12.25 pounds. 

It should be noted here that the weight of air theoretically required 
to burn the hydrogen has been added to the weight actually required 
to burn the carbon as indicated by the flue-gas analysis. While this is 
not exactly correct the error is within the accuracy of the average flue- 
gas analysis. 

It has been previously shown that the coal in question requires 
8.92 pounds of air for theoretical combustion, hence the percentage of 
air excess is 

1AA 11.91- 8.92 QQ _ 

100 o~no = 33.5 per cent. 

The following method, though perhaps not as readily understood as 
the one involving the densities of the several gases, is more expeditious. 
It is based on the principle that the weight of an elementary gas (as 
2 or N 2 ), referred to hydrogen as unity, is proportional to its atomic 



40 STEAM POWER PLANT ENGINEERING 

weight, and that the vapor density of a compound gas (as C0 2 or CO) 
is proportional to one-half its molecular weight. 

Thus the vapor density of C0 2 = J (12 + 2 X 16) = 22, 
and the vapor density of CO = ^ (12 + 16) = 14. 

For the example given above: 

Weight of C0 2 referred to H as unity = 14 X 22 = 308 
Weight of CO referred to H as unity = 0.5 X 14 = 7 
Weight of 2 referred to H as unity = 6 X 16 = 96 

Total 411 

Since C0 2 consists of T 3 T of carbon, and CO, f of carbon, 
Weight of carbon in C0 2 = T 3 T X 308 = 84 
Weight of carbon in CO = f X 7=3 

Total 87 

The weight of oxygen is 411 — 87 = 324. 

The weight of oxygen per pound of carbon = 3 -g^ = 3.72 pounds. 
The weight of air per pound of carbon = 3.72 -f- 0.23 = 16.2 pounds, 
as found by the previous calculations. 

The above method may be expressed algebraically, 
_ 2(C0 2 + 0) + C0 
A2 " 5 ' 8 C0 2 + CO ' (8) 

in which 

A 2 = weight of dry air per pound of carbon. 

C0 2 , CO, and O = percentages by volume of the carbon dioxide, 
carbon monoxide, and oxygen in the flue gas. 

The following formula is based upon the same principle as equation 
(8). (Kent, " Steam Boiler Economy," First Edition, p. 33.) 

11C0 2 + 8Q + 7(C0 + N) 
3 3 (C0 2 + CO) 

in which 

A 3 = weight of dry gas per pound of carbon, other notations as in (8). 

The 7 N in equation (9) represents the N supplied by the air. Since 
the N supplied by the air is 77 per cent of the weight of the air, we 
have 

7N . Q77- 3 - Q4N 

4 3(C0 2 + CO) * U ' 77 C0 2 + CO' (1U; 

in which 

A 4 = the weight of dry air supplied per pound of carbon. 
N, CO, C0 2 = percentages by volume of nitrogen, carbon monoxide, 
and carbon dioxide in the flue gas. 



(9) 



FUELS AND COMBUSTION 



41 



The excess of air supplied per pound of fuel may be conveniently 
determined from the relationship, 

Air actually required N 

Air theoretically supplied 



(11) 



N- 3.782 

N and are respectively, by volume, the proportional parts of the 
nitrogen and oxygen in the flue gas. The free oxygen is due to the air 
supplied and not used. This oxygen was accompanied by 3.782 times 
its volume of nitrogen. (N — 3.782 0) represents the nitrogen con- 
tent in the air actually required for combustion. Hence, N -r- (N — 
3.782 0) is the ratio of the air supplied to that required. 

The various methods discussed above for determining the weight of 
air supplied give practically identical results for pure carbon, but vary 
considerably for fuels containing hydrogen. For fuels containing 
hydrogen and combustible elements other than carbon the amount of 
air necessary to oxidize them should be added to the carbon require- 
ments in determining the total air supply. 



TABLE 9. 

RATIO OF TOTAL AIR SUPPLIED TO THAT THEORETICALLY REQUIRED 
FOR VARIOUS ANALYSES OF FLUE GASES. 

N 



Ratio = 



N- 3.7820 





N=79. 


N = 79.5. 


N=80. 


N=80.5. 


N=81. 


N=81.5. 


N=82.- 


co 2 +co. 


CO2+CO 


CO2 + CO 


CO2+CO 


CO2 + CO 


CO2+CO 


CO2+CO 


co 2 +co 




+ 0=21. 


+ 0=20.5. 


+ O=20. 


+0=19.5. 


+ 0=19. 


+0=18.5. 


+ 0=18. 


21 


1.02 
1.05 
1.11 
1.17 














20 


i ! 66 

1.08 
1.14 


r.oo 

1.05 
1.10 










19 


i ! 62 

1.08 


i ] 66 

1.05 






18 


i ! 62 


Too" 


17 


1.24 


1.20 


1.17 


1.13 


1.10 


1.07 


1.05 


16 


1.32 


1.27 


1.23 


1.20 


1.16 


1.13 


1.10 


15 


1.40 


1.35 


1.31 


1.27 


1.23 


1.19 


1.16 


14 


1.51 


1.45 


1.39 


1.35 


1.30 


1.26 


1.23 


13 


1.62 


1.55 


1.50 


1.44 


1.39 


1.34 


1.30 


12 


1.76 


1.68 


1.61 


1.54 


1.49 


1.43 


1.38 . 


11 


1.92 


1.82 


1.74 


1.66 


1.60 


1.53 


1.48 


10 


2.11 


2.00 


1.90 


1.81 


1.72 


1.65 


1.59 


9 


2.35 


2.21 


2.08 


1.97 


1.88 


1.79 


1.71 


8 


2.65 


2.47 


2.31 


2.18 


2.06 


1.95 


1.86 


7 


3.03 


2.80 


2.59 


2.44 


2.27 


2.14 


2.03 


6 


3.55 


3.22 


2.96 


2.74 


2.54 


2.38 


2.24 


5 


4.27 


3.81 


3.44 


3.14 


2.89 


2.68 


2.50 


4 


5.37 


4.65 


4.11 


3.68 


3.34 


3.05 


2.83 


3 


7.23 


5.97 


5.10 


4.45 


3.96 


3.56 


3.25 


2 


11.06 


8.34 


6.71 


5.63 


4.85 


4.27 


3.82 


1 


23.51 


13.83 


9.83 


7.64 


6.27 


6.12 


4.64 



The weight of air supplied per pound of carbon in the fuel may be 
roughly determined by the percentage of C0 2 in the flue gas. Thus 



42 



STEAM POWER PLANT ENGINEERING 



for the complete oxidation of pure carbon without air excess, the re- 
sulting flue gases should consist of carbon dioxide and nitrogen only, 
and in the ratio by volume of 21 to 79; therefore 21 per cent of C0 2 in 
the flue gas is indicative of complete combustion and theoretical air 
supply. In other words, the ratio by volume of C0 2 to N after com- 
plete combustion is practically the same as the ratio of the oxygen to 
the nitrogen in the air before combustion. This applies only to the 
combustion of pure carbon. For fuels high in volatile matter the per 
cent of C0 2 in the flue gas for complete combustion with theoretical 
air requirements is less than 21 per cent, since the oxygen which com- 
bines with the hydrogen to form H 2 does not appear in the sample of 
flue gas as ordinarily tested; thus for heavy crude oil the correspond- 
ing maximum content of C0 2 is approximately 16 per cent. 

Table 10 gives the weight of air per pound of carbon for different 
percentages of C0 2 in the flue gas. For fuels rich in volatile matter 
these figures may be in considerable error. The per cent of C0 2 in 
the flue gas, however, is an excellent index to the efficiency of com- 
bustion for any fuel. See, also, Table 19. 

In coal-burning practice, from 15 to 16 per cent of C0 2 is all that can 
be expected under the best conditions, with an average between 10 per 
cent and 12 per cent. Anything less than 10 per cent shows an ex- 
cessive amount of air supplied. Traveling grates unless carefully 
operated are apt to show as low as 5 per cent of C0 2 . 

TABLE 10. 

WEIGHT OF AIR PER POUND OF CARBON AS INDICATED BY THE PERCENTAGE OF 

C0 2 IN THE FLUE GAS. 



Per Cent of 


Pounds of 


Per Cent of 


Pounds of 


Per Cent of 


Pounds of 


co 2 . 


Air. 


C0 2 . 


Air. 


C0 2 . 


Air. 


21 


12 


14 


18 


7 


36 


20 


12.6 


13 


19.4 


6 


42 


19 


13.3 


12 


21 


5 


50.5 


18 


14 


11 


22.9 


4 


63 


17 


14.8 


10 


25.2 


3 


84 


16 


15.7 


9 


28 


2 


126 


15 


16.8 


8 


31.5 


1 


210 



The Importance of CO2, as an Index to Combustion and in Connection with Flue 
Gas Temperature, to Boiler Efficiency: Trans. A.S.M.E., 32-1215. Flue Gas Analysis 
and Calculations: Power, Aug. 9, 1910; Eng. Review, Aug., 1910. Real Relation 
of C0 2 to Chimney Losses: Power, Dec. 7, 1909, p. 969. Sampling and Analysis of 
Furnace Gas: Power, Aug. 22, 1911, p. 282. 

See also paragraphs 452-456. 



FUELS AND COMBUSTION 43 

21. Temperature of Combustion. — The actual temperature incident 
to the combustion of a fuel is most satisfactorily determined by means 
of a suitable thermometer or pyrometer. The theoretical temperature 
of combustion may be calculated from the simple relationship 

fe-A-n d2) 

in which 

ti = final temperature of the products of combustion, degrees F. 
h = low calorific value of the fuel, B.t.u. per pound, 
s = mean specific heat of the products of combustion. 
w = weight of the products of combustion, pounds per pound of fuel. 
t = initial temperature of the fuel and air supply, degrees F. 

Thus, in the combustion of one pound of carbon with theoretical air 
requirements, initial temperature 62 degrees F., the maximum theo- 
retical temperature will be 
14 ^40 
k = X2.58X0.29 + 62 = 400 ° degreeS F - ( a PP rox ')- 

No such temperature has ever been obtained in practice from the 
combustion of carbon in air. The discrepancy between actual and 
calculated results is attributed to (1) difficulty of effecting complete 
combustion with theoretical air supply, (2) radiation losses, (3) error in 
the assumed value of the mean specific heat at this high temperature, 
and (4) uncertainty of the proportion of the calorific value of the fuel 
available, at this high temperature, for increasing the temperature of 
the products of combustion. 

An inspection of equation (12) will show that the greater the weight 
of the products of combustion for a given weight of fuel, the lower 
will be the temperature of combustion. Evidently, for maximum tem- 
perature the weight of air supplied per pound of fuel should be kept 
as low as possible, consistent with complete combustion. A perfect 
union of fuel and air in theoretical proportions is almost impossible, 
and to insure complete combustion an excess of air is necessary. The 
influence of air dilution on temperature of combustion is best illustrated 
by a practical example : 

Required the theoretical temperature of combustion of carbon in air 
if 50 per cent air excess is necessary for complete combustion. Since 
the complete oxidation of one pound of carbon requires 11.58 pounds 
of air, the weight of the products of combustion will be 11.58 + 0.5 X 
11.58 + 1 = 18.37 pounds and the final increase in temperature will be 

14 540 
k = 1837 ' Q27 = 3000 degrees F. (approx.). 



44 



STEAM POWER PLANT ENGINEERING 



The values of the mean specific heat (s = 0.29 and s = 0.27) used in 
the preceding computations are based upon the investigations of Pier, 
Holburn and Henning, and Langen, as compiled by Prof. G. B. Upton of 
Cornell University.* Upton recommends the following equations in 
this connection: 



ForN 2 andCO,s 

2 , s 

H 2 , s 

Air, s 

C0 2 ,s 



0.243 + 0.000019 1, (13) 

0.216 + 0.000014/, (14) 

3.369 + 0.00055 1, (15) 

0.237 + 0.000019 *, (16) 

0.2+75Xl0-^-2lXl0- 9 * 2 +2.2Xl0- 12 * 3 , (16a) 



H 2 0,s = 0.452+7.4 Xl0-n+92.6Xl0-n 2 -20.6Xl0- 12 / 3 , (16b) 



in which 



s = mean specific heat at constant atmospheric pres- 
sure and temperature range degrees C. to t 
degrees C. 

t = maximum temperature, degrees C. 

Between 1000 degrees C. and 1500 degrees C. the results are un- 
certain and dependence can be placed in only the first two significant 
figures in the decimal. Beyond 1500 degrees C. the results are purely 
conjectural since experiments have not been made at these high tem- 
peratures. 



*.28 



O .27 
45.86 



100 





















































































































































































































































































































































































































































































































_a 


ad 


C 
































































































& 






































































































— 




















c 0<! 




\s 










































































s 











































































* 


y 










































































y\ 


























































































































































































o 














































































\'i 








































-" 















































































































































































































































































































































































































2* 

• C3 S 

.61 © 
.59 « 
.57 a 
.55 £ 
•53 '§ 
•51 £ 

.47 J 
.45 S 



4.5 
4.4 c* 

4.3 W 
4:2 *S 
4.1 § 

4.0 <£ 
3.9 «l 
3.8 I 

it t 

3.4 s 
3.3 



300 400 500 600 700 800 900 1000 1100 1200 1300 1400 1500 1600 1700 1800 1900 2000 
Temperature, Beg. C. 
32 212 392 572 152 932 1112 1292 1472 1652 1832 2012 2192 2372 2552 2732 2912 3092 3272 3452 3638 
Temperature, J)eg-. J?. 
Fig. 7. Mean Specific Heat at Constant Pressure. 

The application of the formulas for the mean specific heats at high 

temperatures to equation (12) necessitates laborious calculations, and 

since the results are only approximate at the best extreme refinement 

in calculation is without purpose. The curves in Fig. 7 are plotted 

* "Experimental Engineering," Carpenter and Diederichs, 1911, p. 865. 



FUELS AND COMBUSTION 45 

from the above equations, and afford a means of approximating the 
mean specific heat without the labor of solving the equations. 

If the mean specific heats, s h s 2 . . . s n , and weights, w h w 2 . . . w n , 
of the constituent gases of a compound are known the mean specific 
heat, s, of the compound may be determined as follows: 

WiSi + w 2 s 2 . • • + w n s n 

s = i i (17) 

Wi + w 2 . . . + w n 

The mean specific heat between any two temperatures, h and t, may 
be determined by substituting (ti + t) for t in above equation for the 
mean specific heat between zero and t degrees. 

22. Heat Losses in Burning Coal. — A boiler in order to entirely 
utilize the heat of combustion of the fuel must be free from radiation 
and leakage losses, the fuel must be completely oxidized and the prod- 
ucts of combustion must be discharged at atmospheric temperature. 
Commercially such conditions are unobtainable, hence complete utili- 
zation of the heat generated is impossible. A boiler which utilizes 
83 per cent of the heat value of the fuel is exceptional and an average 
figure for very good practice is not far from 75 per cent, The various 
losses may be summed up as follows : 

1. Loss in the dry chimney gases. 

2. Loss due to incomplete combustion. 

3. Loss of fuel through the grate. 

4. Superheating the hygroscopic moisture in the air. 

5. Moisture in the fuel. 

6. Loss due to the presence of hydrogen in the fuel. 

7. Unburned fuel carried beyond the combustion chamber in the 
form of soot or smoke. 

8. Radiation and minor losses. 

Some of these losses are preventable. Others are inherent and can- 
not be avoided. 

23. Loss in the Dry Chimney Gases. — This loss depends upon the 
type and proportion of the boiler and setting and upon the rate of 
driving. It is usually the greatest of all the losses. The heat carried 
away may be expressed: 

h = W (t c - t)c. (18) 

in which 

hi = B.t.u. lost per pound of fuel, 

W = weight of dry chimney gases per pound of fueL 

t c = temperature of the escaping gases, degrees F. 

t = temperature of the air entering the furnace. 

c = mean specific heat of the dry gases. (This may be taken as 
0.24 for most purposes.) 



46 



STEAM POWER PLANT ENGINEERING 



It will be noted that the magnitude of this loss depends chiefly upon 
the air dilution and the temperature at which the gases are discharged. 
Flue temperatures less than 450 degrees F. are seldom experienced 

TABLE 11. 

HEAT CARRIED AWAY BY THE DRY CHIMNEY GASES PER POUND OP 

COMBUSTIBLE. 



"■3 
to 

3 
XI 

E 
o 
o 

"S 

■a 

3 
O 

fk 

u 

p. 

v-t 

o 

W 
CI 
§ 




Temperature of Chimney Gases. Deg. Fahr. 


300° 


350° 


400° 


450° 


500° 


550° 


600° 


650° 


12 

* 


750 

5.2 


905 

6.2 


1060 

7.3 


1216 

8.7 


1370 

9.5 


1528 

10.5 


1683 

11.6 


1840 

12.7 


lS 


865 

6 


1112 

7.6 


1305 

9.1 


1498 

10.3 


1679 

11.6 


1880 
13.0 


2072 

14.3 


2262 

15.6 


18 


1004 

7.2 


1321 

9.1 


1550 

10.7 


1778 
12.2 


2010 

13.9 


2235 

15.4 


2460 

17 


2692 

17.9 


21 


1266 

8.7 


1530 

10.5 


1785 

12.3 


2060 

14.2 


2320 

16 


2582 

17.8 


2846 

19.5 


3118 

21 


24 

27 


1440 

9.9 


1740 

12 


2040 

14 


2340 

16.1 


2640 

18.2 


2940 

20.3 


3240 

22.4 


3540 

24.4 


1611 

11.1 


1950 

13.5 


2281 

15.7 


2620 

18.1 


2958 

20.4 


3291 

22.7 


3628 

25 


3962 

27.4 


30 
33 
36 
39 


1785 

12.4 


2160 

14.9 


2530 

17.4 


2900 

20 


3270 

22.6 


3641 

25 


4016 

27.8 


4396 

30.4 


1957 

13.5 


2362 

16.3 


2779 

19.2 


3180 

22 


3589 

24.7 


4000 

27.6 


4405 

30.5 


4820 

33.2 


2130 

14.7 


2579 

17.8 


3020 

20.8 


3461 

23.9 


3910 

27 


4350 

30 


4798 

33 


5290 

36.6 


2300 

15.9 


2781 
19.2 


3261 

22.5 


3743 

25.8 


4220 

29.2 


4700 

32.4 


5180 

35.7 


5670 

39 


42 


2479 
17.1 


2999 

20.6 


3508 

24.7 


4023 

27.7 


4540 

31.3 


5052 

34.8 


5570 

39.4 


6100 

42 



* Theoretical requirement. 

Large type gives the loss in B.t.u. per pound of combustible. 
Small type gives the per cent loss, assuming a calorific value of 14,540 B.t.u. 
per pound of combustible. 

except in connection with economizers, and the air dilution is ordinarily 
in excess of 50 per cent of theoretical requirements, hence the loss from 
this cause may range from 8 per cent to 40 per cent of the total heat 



FUELS AND COMBUSTION 



47 



generated. In excellent practice it is not far from 12 per cent with a 
general average of from 20 to 25 per cent. In exceptional cases a loss 
from this cause as low as 9 per cent has been recorded. (Jour. A.S.M.E., 
Nov., 1911, p. 1463.) 

Table 11 indicates the magnitude of the losses for different chimney 
temperatures and weights of air per pound of combustible. 

24. . Loss Due to Incomplete Combustion. — If the volatile gases are not 
completely oxidized, as when the air supply is insufficient or the mix- 
ture of air and gases is not thorough, some of the carbon may escape 
as CO. Some of the hydrocarbons may also pass through the furnace 
without being burned. (See Table 12.) The presence of even a small 
amount of CO in the flue gas is indicative of a very appreciable loss, 
as will be seen from Table 14. Carbon monoxide is a colorless gas 
and its presence in the chimney gases cannot be detected by the fire- 
man, consequently the absence of smoke is not an infallible guide for 
perfect combustion. Since the heat of combustion of C to CO is but 
4380 B.t.u. against 14,540 B.t.u. for complete combustion of C to C0 2 
this loss may be expressed 

(14,540 - 4380) CO 



h* = C 



C 



C0 2 + CO 
10,160 CO 



(19) 



in which 



C0 2 + CO 



h 2 = the loss in B.t.u. per pound of fuel, 
C = proportional part of carbon in the fuel, 
C0 2 and CO are percentages by volume of the flue gases. 



TABLE 12. 

ANALYSIS OF CHIMNEY GASES. 
(Report of Committee for Testing Smoke-preventing Appliances, Manchester, England, 1905.) 



Boiler. 


Smoky. 


Clear. 


co 2 


o 2 


CO 


CH 4 


H 2 


N 2 


co 2 


o 2 


CO 


CH 4 


H 2 


N 2 




111.00 
\10. 65 


6.90 
6.45 


0.90 
2.15 






81.20 
80.75 






































17.00 
\9.00 


13.50 
9.75 










79.50 








81.25 


No. 2, hand fired 


i6.25 
13.25 

10.95 

8.75 


8.60 
3.50 

1.30 

7.00 


.50 
.05 

3.00 

3.25 



0.25 

.70 

.40 





3.23 

1.00 


80.65 
82.95 

80.82 

79.60 








No. 3, hand fired 














No. 4, fire under caustic 














No. 5, split bridge, hand 














No. 6, with smoke-pre- 


7.25 
7.15 
8.15 


12.00 
12.15 
11.10 


















80.75 


No. 7, with smoke-pre- 














80.70 


No. 8, with smoke-pre- 
vention device 














80.75 



















48 



STEAM POWER PLANT ENGINEERING 



TABLE 13. 

RELATION OF CO AND COMBUSTION-CHAMBER TEMPERATURES. 

(U. S. Geological Survey). 





Per Cent of Black Smoke. 







OtolO 


10 to 20 


20 to 30 


30 to 40 


40 to 50 


50 to 60 


Number of tests 


37 

0.05 

9.14 


18 
7.1 
0.11 

10.60 


56 
15.5 
0.11 

9.46 


51 

24.7 

0.14 

10.93 


36 
34.7 
0.21 

11.41 


17 
43.1 
0.33 

13.41 


4 


Average per cent of smoke 


52.9 


Average per cent of CO in flue gases 

Average per cent unaccounted for in heat 
balance 


0.35 
13.34 






Number of tests * 


26 
2180 


16 ' 
2215 


48 
2357 


45 

2415 


32 
2450 


17 
2465 


4 


Average combustion-chamber temperature 

(°F.)... 


2617 







* Temperatures in combustion chamber were not determined on all tests. 

This loss may be reduced to a negligible quantity in a properly designed 
and carefully operated furnace. In fact the loss from this cause is 
often exaggerated and seldom exceeds 2 per cent of the total heat value 
of the fuel except during the fev^ moments following the replenishing 
of a burned-down fire with fresh fuel or when the supply of air is checked 
to meet a sudden reduction in load. In improperly designed furnaces 
in which the volatile gases are brought into contact with the cooler 
boiler surfaqe before combustion is complete, the carbon monoxide may 



15 
14 
13 
















































































































- 
















































































c 


Jji_ 






























X2 

11 

10 

9 

8 

7 

6 
5 
<\ 
















































































Relation of Gas Composition in Rear 

Combustion Chamber To Temperature 

at Same Place 






























































































































































































































L^2 




















































































3 
3 

1 
o 








































































































































C 


0__ 























































































,1900 2000 



2700 



2800 



Fig. 8. 



2100 2200 2300 2400 2500 2& 

Combustion Chamber Temperature.Deg.Fah. 
Relation of Gas Composition in Combustion Chamber to Temperature. 



be reduced in temperature below its ignition point and consequently 
will fail to combine with the oxygen. In such a case the loss may 
prove to be a serious one. Fig. 8 shows the relation between the com- 
position of the products of combustion in the rear combustion chamber 
of a 250-horse-power Heine boiler, hand-fired, and the temperature at 



FUELS AND COMBUSTION 



49 



the same place. (For an extended discussion of this subject see Jour. 
West. Soc. Engrs., June, 1907, p. 285.) 

25. Loss of Fuel through Grate. — The refuse from a fuel is that 
portion which falls into the pit in the form of ashes, unburned or partially 



TABLE 14. 

LOSS DUE TO INCOMPLETE COMBUSTION OF CARBON TO CARBON 

MONOXIDE. 



6 

s 

"o 
> 
>> 
& 

a 

xi 

a 

O 
O 

"o 

S3 
03 
D 
M 
V 
P4 




Per Cent of C0 2 in the Flue Gas by Volume. 


6 


s 


10 


12 


14 


16 


0.2 


328 

2.2 


248 
1.7 


199 

1.3 


168 

1.1 


144 

1 


126 

0.8 


0.4 


635 

4.3 


484 

3.3 


390 

2.6 


327 

2.2 


282 
1.9 


248 
1.7 


0.6 


925 

6.3 


709 

4.8 


575 

3.9 


474 

3.2 


417 

2.8 


367 

2.5 


0.8 


1192 

8.1 


923 

6.3 


750 

5.1 


635 

4.3 


549 

3.7 


495 

3.4 


1.0 


1494 

10.2 


1128 

7.7 


923 

6.3 


780 
5.3 


676 

4.6 


59G 

4.1 


1.2 


1690 

11.5 


1321 

9 


1085 

7.4 


923 

6.3 


801 
5.4 


708 
4.8 


1.4 


1920 

13.1 


1512 

10.3 


1248 

8.5 


1061 

7.2 


924 

6.3 


819 

5.6 


1.6 


2104 

14.3 


1693 

11.5 


1400 

9.5- 


1193 

8.1 


1040 

7.1 


924 

6 3 


1.8 


2340 

16 


1865 

12.7 


1549 

10.5 


1321 

9.0 


1151 

7.8 


1025 

7 


2.0 


2537 

17.2 


2030 

13.8 


1690 

11.5 


1450 

9.9 


1270 

8.6 


1129 

7.7 



Large type gives the loss in B.t.u. per pound of carbon. Small type gives the 
per cent loss, assuming a calorific value of 14,540 B.t.u. per pound of carbon. 

burned fuel, and cinders. The loss from this cause depends upon the 
size of the fuel, the width of opening in the grate bars, and the type of 
grate. Coal which necessitates frequent slicing is apt to give greater 
loss than a free-burning coal. Under good conditions of operation it 
ought not to exceed 4 per cent of the total heat value of the fuel. In 



50 STEAM POWER PLANT ENGINEERING 

traveling grates in which a large percentage of the fine fuel falls through 
the front end of the grate a special hopper is ordinarily installed in the 
ash pit which reclaims most of it. (See Fig. 102.) 

If h c = calorific value of combustible in the dry refuse, 
y = percentage of combustible in the dry refuse, 
a = percentage of ash in the coal as fired, 
hz = heat loss in the refuse, B.t.u. per pound of coal as fired, 

h c I ya 



h - 1 



w^> 



Since the heat balance is only an approximation at the best the calorific 
value of the combustible in the refuse may be taken as that of the 
combustible fired. 

26. Superheating the Moisture in the Air. — The loss due to this 
cause is a minor one, though on hot, humid days it may be appreciable. 
This loss may be expressed 

h* = Mc (t c - t), (21) 

in which 

h A = B.t.u. lost per pound of fuel, 
M = weight of moisture introduced with the air per pound of fuel, 

c = mean specific heat of water vapor, t to t c degrees F., 

t = temperature of air entering the furnace, degrees F., 
t c = temperature of chimney gases, degrees F., 
M = zwvA, (22) 

in which 

z = relative humidity (see paragraph 271), 

w = weight of 1 cubic foot of water vapor at t degrees F. (this may 
be taken directly from steam tables), 

v = volume of 1 pound of dry air at t degrees F., cubic feet, 
A = weight of dry air supplied per pound of fuel. 

27. Moisture in the Fuel. — Moisture in the fuel represents an 
appreciable loss in economy if present in large quantities, since the 
heat necessary to evaporate it into superheated steam at chimney 
temperature is lost. Firemen occasionally wet the coal to assist coking 
or to reduce the dust, but moisture thus added necessarily reduces 
the theoretical furnace efficiency. Under certain conditions wet coal 
may give a higher evaporation than dry coal. (See paragraph 102.) 

The loss due to evaporating the moisture may be expressed 

h b = W (1150.4 - a (t - 32) + c't.)* (23) 

= W (1182.4 - t + c%), (24) 

* For all purposes C\ may be taken as unity and in the following calculations 
and equations it has been considered as such. 



FUELS AND COMBUSTION 51 

in which 

h 5 = B.t.u. lost per pound of fuel, 
1150.4 = total heat of one pound of saturated steam above 32 de- 
grees F. at atmospheric pressure, 
Ci = mean specific heat of water, 32 to t degrees F., 
W = weight of moisture per pound of fuel, 
' c' = mean specific heat of water vapor, 212 to t c degrees* F., 
t = temperature of the fuel, 
t c as in equation (21), 
t 8 = degree of superheat = (t c — 212). 

Carpenter and Diederichs ("Experimental Engineering," 1911, p. 467) 
give this heat loss as 

h 6 = W (1090.7 + 0.455 t c - t). (25) 

The difference in results between the two formulas is less than T 3 o of 
one per cent for ordinary conditions of practice. 

28. Loss due to the Presence of Hydrogen in the Fuel. — The hydro- 
gen in any fuel which is not rendered inert by oxygen burns to water 
and in so doing liberates 61,950 B.t.u. per pound. All of this heat is 
not available for producing steam in the boiler, since the water formed 
by combustion is discharged with the flue gases as superheated steam 
at chimney temperature. This loss is equal to 

h 6 = 9H (1182.4 - t + c%), (26) 

in which h Q = B.t.u lost per pound of fuel, 

H = weight of hydrogen per pound of fuel. 

All other notations as in equations (24) and (25). 

With anthracite coal this loss is approximately 2.5 per cent of the 
total heat value of the combustible and with bituminous coal it runs as 
high as 4.5 per cent. 

29. Loss due to Smoke. — Visible smoke consists of carbon in a 
flocculent state and ash mixed with the products of combustion. It is 
seldom evident in connection with anthracite coal and is generally 
associated with bituminous fuel. A smoky chimney does not neces- 
sarily indicate an inefficient furnace, since the losses due to visible 
smoke generation seldom exceed 2 per cent ; as a matter of fact, a smoky 
chimney may be much more economical than one which is smokeless. 
That is to say, a furnace operating with minimum air supply may 
cause dense clouds of smoke and still give a higher evaporation than 
one made smokeless by a very large excess of air. There will be some 
loss due to carbon monoxide, unburned hydrocarbons and soot in the 
former case, but this may be more than offset by the excessive losses 



52 



STEAM POWER PLANT ENGINEERING 



caused by the heat carried away in the chimney gases in the latter. 
The amount of combustible in the soot and cinders deposited on the 
tubes and in various parts of the setting seldom exceeds two per cent 
of the calorific value of the fuel. 

Smoke has become such a public nuisance, particularly in the larger 
cities, that special ordinances prohibiting its production have been 
enacted and violators are subject to heavy fines. Effective enforce- 
ment of these ordinances renders smoke production very costly and 
the problem of smokeless combustion becomes a momentous one. 

The subject of smoke prevention and smoke-prevention devices is 
discussed at some length in Chapter V. 

30. Radiation and Minor Losses. — These losses are usually deter- 
mined by difference. That is, the difference between the heat repre- 
sented in the steam and the losses just mentioned is charged to radiation, 
leakage, and unaccounted for. Summing up the various losses we have 
as a rough approximation 



Poor 
Practice. 
Per Cent. 



Heat given to steam 

Loss in chimney gases 

Loss due to carbon burning to CO 

Loss of fuel through grate 

Loss due to moisture in coal, moisture 

in air, and hydrogen in fuel 

Smoke, soot, etc 

Radiation and minor losses 



Excellent 


Good 


Average 


Practice. 


Practice. 


Practice. 


Per Cent. 


Per Cent. 


Per Cent. 


80 


70 


60 


12 


18 


24 





1 


2 


0.5 


1 


2 


3.0 


3 


3 





0.5 


1 


4.5 


6.5 


8 



50 
30 

3 

3 

3.5 
1.5 
9 



The above heat-loss distribution refers specifically to boilers in con- 
tinuous operation. In many situations, particularly in large central 
stations, a considerable portion of the boiler equipment is held in re- 
serve for peak loads and as a consequence the boilers are in actual oper- 
ation but a short period during the day. The coal burned in banking 
the fires in order to hold the boilers in readiness is charged to standby 
losses. These losses include the coal required to start up cold boilers, 
the heat discharged in "blowing off," the fuel lost in cleaning fires and 
in shutting down. The magnitude of standby losses depends upon 
the size and character of the boiler equipment and the condition of 
operation, and may range from 5 per cent to 15 per cent or more of the 
total heat generated (yearly basis). 

Example: Calculate the various heat losses from the following data: 
Heat absorbed by the boiler, 76 per cent of the calorific value of the 
coal as fired. 






FUELS AND COMBUSTION 



53 



Analysis of coal as fired : 

Per Cent. 

Carbon 65 

Oxygen 8 

Hydrogen 5 



Per Cent. 

Ash and sulphur 13 

Water 8 

Nitrogen 1 



Calorific value as fired, 11,850 B.t.u. 



Flue-gas analysis: 



C0 2 . 
2 . 



Per Cent. Per cent. 

. 14 CO 0.5 

6 N 79.5 (by difference). 



Temperature of air entering furnace, 70 degrees F.; temperature of 
flue gases, 470 degrees F.; temperature of the steam in the boiler, 340 
degrees F.; relative humidity of air entering furnace, 80 per cent; 
combustible in the dry refuse, 20 per cent. 

The heat distribution may be referred to the coal as fired, dry coal 
or combustible. In this problem it is referred to the coal as fired. 

DISTRIBUTION OF ACTUAL HEAT LOSSES PER POUND OF COAL AS FIRED. 

(Calorific Value of Coal as Fired, 11,850 B.t.u.) 





B.t.u. 


Per Cent. 





Heat absorbed by the boiler, 0.76x11,850 


9006 
1176 

227 

487 

27 

98 

554 
275 


76.00 


1. 


Loss in the dry chimney gases, 

hi = 12.25* (470 - 70) 0.24 (equation 18) 


9.92 


2. 


Loss due to incomplete combustion, 

, n „ 10,160 X 0.05 , ,. 1ft . 

ho = 0.65 — (eauation 19) : . 


1.91 


3. 

4. 
5. 
6. 

7 


Loss in combustible in refuse, 

11,850 -r- 0.79 / 20 X 13 \ , om 

h= 100. (l00-20J (eqUatlOn20) 

Loss due to moisture in the air, 

A 4 =0.8x0.00115x 13.3X 11.91* X0.46 (470-70) (eq.22)... 
Loss due to moisture in the coal, 

h = 0.08 [(1182.4 - 70) + 0.46 (470 - 212)] (equation 24) . . 
Loss due to hydrogen in the fuel, 

h =9 X 0.05 [(1182.4 - 70 + 0.46 (470 -212)] (eq. 26) 

Radiation and unaccounted for (by difference) 


4.11 

0.23 

0.83 

4.67 
2.33 




Total 






11,850 


100.00 









* See third example, paragraph 20. 

The inherent or unpreventable losses may be summarized as follows : 

1. Heat absorbed by the theoretical weight of dry chimney gases in 
being heated from boiler room to boiler steam temperature. 

2. Heat required to evaporate and superheat the moisture in the 
fuel from boiler room to boiler steam temperature. 

3. Heat required to evaporate and superheat the H 2 formed by 
the combustion of hydrogen in the fuel from boiler room to boiler steam 
temperature. 



54 



STEAM POWER PLANT ENGINEERING 



4. Heat required to superheat the moisture in the air (theoretical 
requirements) from boiler room to boiler steam temperature. 
For the preceding problem: 

DISTRIBUTION OF INHERENT HEAT LOSSES PER POUND OF COAL 

AS FIRED. 





B.t.u. 


Per Cent. 


1. Inherent loss in the dry chimney gas, 

9.26* X (340-70) 0.24 


600.0 
93.7 

427.0 

13.5 
10,715.8 


5.06 


2. Inherent loss due to moisture in coal, 

0.08 [1182.4-70+46 (340-212)] 


0.79 


3. Inherent loss due to H 2 formed by the combustion of 
hydrogen, 
9X0.05 [1182.4-70+0.46 (340-212)] 


3.60 


4. Inherent loss due to " humidity " of the air, 

0.8 X 0.00115 X 13.3 X 8.92* X 0.46 (340-70) 


0.11 


5. Heat absorbed by ideal boiler (by difference) 


90.44 






Total 


11,850.0 


100.00 







* See first example, paragraph 20. 

31. Size of Coal — Bituminous. — Coal is usually marketed in different 
sizes, ranging from lump coal to screenings. The latter furnish by far 
the greater part of the stoker fuel used. For maximum efficiency 
coal should be uniform in size. With hand-fired furnaces there is 
usually no limit to its fineness and larger sizes can be used than with 
stokers. As a rule the percentage of ash increases as the size of coal 
decreases. This is due to the fact that all of the fine foreign matter 
separated from larger coal, or which comes from the roof or the floor of 
the mine, naturally finds its way into the smaller coal. The size best 
adapted for a given case is dependent upon the intensity of draft, 
kind of stoker or grate, and the method of firing, and its proper selec- 
tion often affords an opportunity to effect considerable economy. 
The influence of the size of screenings on the capacity and efficiency of 
a boiler in a specific case is illustrated in Fig. 9. The curves are 
plotted from a series of tests conducted with Illinois screenings on a 
500-*horse-power B. & W. boiler, equipped with chain grates, at the 
power house of the Chicago Edison Company. 

Influence of Thickness of Fire. — See paragraph 82. 
Size of Coal: Some Characteristics of Coal as affecting Performances with Steam 
Boilers: Jour. West. Soc. Engrs., Oct., 1906, p. 528. 

32. Washed Coal. — Coal is washed for the purpose of separating 
from it such impurities as slate, sulphur, bone coal, and ash. All 
of these impurities show themselves in the ash when the coal is 
burned. Screenings contain anywhere from 5 per cent to 25 per cent 



FUELS AND COMBUSTION 



55 



of ash and from 1 per cent to 4 per cent of sulphur. Washing eliminates 
about 50 per cent of the ash and some of the sulphur. Table 15 gives 
some idea of the effects of washing upon a number of grades of coal. 

















































































































oi 


f 






























rf 


V 
































V 




























80 




A 






























z_ 






























/ 












iTiCSs 






















/ 








~®fru* 
































i, 
























































Influence of-Size of Coal on the Capacity 
and Efficiency of a B.& W.Boiler, Chain Grate 
Keating Surface 5000 Sq.Ft. 
Superheating Surface 1000 Sq.Ft. 
























40 






























































\ 


















&§£X 


Coo. 


gc£^ 












\ 


<; 


» — 






























X 


20 



































































1000 



35 



306 
u 
<o 

25 P* 

200 20 I 

d 



10 



1.25 1.00 0.75 0.50 0.25 

Size of Coal in Inches 

Fig. 9. Influence of Size of Coal on Boiler Capacity and Efficiency. 

The evaporative power of the combustible is practically unaffected 
by washing and the greater part of the water taken up by the coal is 
removed by thorough drainage. Many coals otherwise worthless as 
steam coals are rendered marketable by washing. Washed coals are 
usually graded as follows: 



Size. 


Screens. 


No. 1 


Over If 


Under 2| 


2 


1* 


If 


3 


3 
4 


1* 


4 


1 
4 


3 

4 


5 




i 



56 



STEAM POWER PLANT ENGINEERING 



Numbers 3 and 4 are excellent sizes for use in connection with stokers 

and No. 5 is well adapted for hand furnaces where smoke prevention is 

essential. 

TABLE 15. 

EFFECT OF WASHING ON BITUMINOUS COALS. 

{Journal W.S.E., December, 1901.) 



Before Washing. 
(Per Cent.) 



Ash. 



Sul- 
phur. 



Fixed 
Carbon. 



After Washing. 
(Per Cent.) 



Ash. 



Sul- 
phur. 



Fixed 
Carbon. 



Belt Mountain, Mont 

Wellington Colliery Co., Van- 
couver Island (new coal) . . . 

Alexandria Coal Co., Crabtree, 
Pa 

DeSoto, 111 

Northwestern Improvement 
Co., Roslyn, Wash 

Luhrig Coal Co., Zaleski, Ohio 

Rocky Ford Coal Co., Red 
Lodge, Mont 

Buckeye Coal and Ry. Co., 
Nelsonville, Ohio 

New Ohio Washed Coal Co., 
Carterville, 111 



18.74 

35.00 

10.60 
18.00 

16.30 
15.80 

25.30 

13.77 

9.48 



3.34 



43.72 
38.00 



1.30 



0.57 
1.90 



44.00 
45.90 



1.05 
0.78 



37.80 
49.04 
55.00 



5.56 

8.90 

6.21 
4.20 

9.70 
8.00 

8.50 

4.30 

4.85 



2.40 



0.61 



0.40 
0.87 



0.89 
0.69 



48.39 
56.90 



57.00 

47.86 
50.90 

47.20 

54.82 

63.00 



33. Purchasing Coal. — Engineers fail to agree as to the specifica- 
tions best suited for the purchase of coal. Some extensive purchasers 
require elaborate analyses and others specify only the size and grade 
of the fuel. Every essential requirement of the purchaser may be 
fulfilled by confining them to the four following characteristics: 

Moisture. 

Ash. 

Size of coal. 

Calorific value of the coal. 

Although moisture is a great and uncertain variable, and the producer 
can exercise no control over this factor, still the purchaser should pro- 
tect himself against excessive moisture by stipulating an amount con- 
sistent with the average inherent moisture in the coal, and proper penalty 
should be fixed for delivery in excess of the amount allowed, a corre- 
sponding bonus being paid for delivery of less than contract amount. 
Considerable attention should be given to the percentage of earthy 
matter contained. The amount of earthy matter usually fixes the 
heating value of the coal, since the heating value of the combustible 



FUELS AND COMBUSTION 



57 



is practically constant. The effect of ash on the heat value of Illinois 
screenings as fired under a B. & W. boiler with chain grate is shown 
in Fig. 10. This value varies with the different types of boilers, grates, 
and furnaces, but is substantially as illustrated. The amount of refuse 



100 


































90 


































































80 




































































































70 


































g 60 
> 
























































\ 










> 

I 50 

t 
























\ 
































\ 


\ 








8 

40 


























\ 
































\ 








30 




























\ 
































\ 






£0 




























\ 








Influence of Ash on Fuel Value of Dry- 
Coal. (Illinois Screenings) 
B.& W. Boiler, Chain Grate. 
Screenings with 12.5 Per Cent" Ash ■ 
taken at 100. 










\ 




10 












\ 
































\ 


























four.S 


».W.I 


.Oct.l 


306 E. 


>42. \ 





10 20 30 40 

Per Cent of Asn'in Dry CoaL 

Fig. 10. Influence of Ash on Fuel Value of Dry Coal. 

in the ash pit is always in excess of the earthy matter as reported by 
analysis. 

The maximum allowable amount of sulphur is sometimes specified, 
since some grades of coal high in sulphur cause considerable clinker- 
ing. But sulphur is not always an indication of a clinker-producing 
ash, and a more rational procedure would be to classify a coal as clinker- 



58 STEAM POWER PLANT ENGINEERING 

ing or non-clinkering according to its behavior in the particular furnace 
in question, irrespective of the amount of sulphur present. An analysis 
of the various constituents of the ash is necessary to determine whether 
or not the sulphur unites with them to produce a fusible slag, and as 
such analyses are usually out of the question on account of the expense 
attached they may well be omitted. 

The heating value of the coal as determined by a sample burned in an 
atmosphere of oxygen does not give its commercial evaporative power, 
since this depends largely upon the composition of the fuel, character 
of grate, and conditions of operation. It serves, however, as a basis 
upon which to determine the efficiency of the furnace. In large plants 
where a number of grades of fuel are available it is customary to con- 
duct a series of tests with the different grades and sizes, and the one 
which evaporates the most water for a given sum of money, other con- 
ditions permitting, is the one usually contracted for. In designing a 
new plant particular attention should be paid to the performance of 
similar plants already in operation, and that fuel and stoker should be 
selected which are found to give the best returns for the money. Where 
smoke prevention is a necessity the smoke factor greatly influences the 
choice of fuel and stoker. (See, also, paragraph 480.) 

The Purchase of Coal: Eng. Mag., Mar., 1911; Jour. A.S.M.E., Mar., 1911; 
Power, Apr. 6, 1909, p. 642. 

The Purchase of Coal by the Government under Specifications: Bureau of Mines, 
Bull. No. 11, 1910; U. S. Geol. Survey, BuUetins No. 339, 1908; No. 378, 1909. 

The Fusing Temperature of Coal Ash: Power, Nov. 28, 1911, p. 802. 

The Clinkering of Coal: Eng. News, Dec. 8, 1910. 

34. Powdered Coal. — The value of powdered coal as a fuel for steam 
boiler plants has long been known, and appliances for pulverizing and 
feeding the coal have been on the market for a number of years. How- 
ever, despite the many advantages of powdered fuel and the apparent 
success of some of the systems of burning it, little progress has been 
made toward its general adoption. 

Some of the advantages obtained in burning powdered coal are: 

(a) Complete combustion and total absence of smoke. The coal 
in the form of dry impalpable dust is induced or forced into the zone 
of combustion, where each minute particle is brought into contact 
with the necessary amount of air and complete oxidation is effected 
without the excess of air which accompanies the firing with lump coal, 
provided the furnace is properly proportioned. With a properly 
designed setting there is complete absence of smoke. 

(b) A cheaper grade of bituminous coal may be burned, since the 
per cent of ash and moisture has little effect on the completeness of 






FUELS AND COMBUSTION 



59 



combustion and the full value of the fuel is more nearly realized than 
with ordinary firing. 

(c) The plant may be rapidly forced above its rated capacity and 
sudden demands for power readily met. 

(d) The labor of firing is reduced to a minimum. 

35. Furnaces for Burning Powdered Coal. — In burning ordinary bulk 
coal the mass of incandescent fuel stores up a quantity of heat to effect 
the distillation and ignition of the volatile matter in the green fuel. 
With pulverized coal a refractory lining is necessary to bring about the 
same result. A large combustion chamber is necessary and the shape 
of furnace and path of flame must be such as to provide a uniform dis- 
tribution of heat over the boiler-heating surface without direct im- 
pingement of flame.. Fire-brick arches and target walls are not to be 
recommended, owing to the rapid destruction of the brickwork under 
the intense heat of combustion. Fig. 11 shows a successful arrange- 
ment of boiler and furnace for burning coal dust. (For detailed de- 
scription of this apparatus, see Power, Feb. 14, 1911, p. 264.) 

36. Types of Powdered-coal Burners. — Powdered-coal burners may 
be grouped into two general classes: 

1 . The dust-feed burner, in which the coal is supplied in the powdered 
form, and 

2. The self-contained burner, in which the coal is crushed, pulver- 
ized, and fed to the furnace simultaneously. 

The dust may be fed into the furnace by 

1. Natural draft, 

2. Mechanical means, or by 

3. Forced draft. 

The following outline gives a classification of a few of the best known 
coal-dust burners: 



Natural Draft 



Forced Draft 



Natural Draft (pinther 
Feed [Wegener 

Brush Feed Schwartzkopff 



Blower Feed 



Compressed Air 
Paddle Wheel 



Dust Feed 



/Cyclone 
\Triumph 

f Eng. and Powdered 
(.Fuel Company 

j Ideal 
1 Blake 



Self-contained 



37. Pinther Apparatus. — Fig. 12 shows a section through a Pinther 
coal-dust feeder, illustrating the principles of the "natural-draft feed" 
type. The powdered coal is placed into hopper B, from which it is 
fed by rollers a, a into the chamber leading to the furnace C. The dust 



60 



STEAM POWER PLANT ENGINEERING 



e.mSoa paaj 




FUELS AND COMBUSTION 



61 



falls in a thin stream and is caught up by the current of air and drawn 
into the furnace as indicated. The furnace is lined with refractory 
material heated to a sufficiently high temperature to ignite the fuel 
and burn it in suspension. The chief drawback to a burner of this 
type is its limited capacity. Any attempt to feed large quantities of 
fuel into the furnace necessitates such a strong current of air as to 
carry the particles of dust beyond the zone of combustion before they 
are completely consumed. Within the limits of its capacity it is an 
efficient and simple apparatus, but is open to the same objection as all 
burners of this type in that it necessitates the storage of powdered 
coal. This apparatus is not much in evidence in boiler plants. 




Fig. 12. Pinther Coal-dust Feeder. 



Fig. 13. Schwartzkopff Coal-dust Feeder. 



38. Schwartzkopff Apparatus. — Fig. 13 shows a section through a 
Schwartzkopff feeder, illustrating the principles of the brush-feed, 
natural-draft system. It is a very simple and practical dust feeder, 
though open to the objection of all systems which require the coal to be 
ground and pulverized in separate machines. The fuel is placed in a 
hopper and its supply to the brush is regulated by the hand screw A 
and the spring plate bottom of the hopper. The brush, consisting of 
a number of flat steel leaves ^V inch by | inch wide, revolves at a high 
speed, 1000 to 1200 r.p.m. and forces the dust into the furnace. The air 
for combustion is admitted either through the grates in the ordinary 



62 



STEAM POWER PLANT ENGINEERING 



way or through the lower chamber of the burner. To prevent the dust 
from bridging in the hopper, a small hammer C is fitted to the brush 
so that it will strike the plate D and agitate the dust. This apparatus 
is meeting with much success in connection with annealing furnaces, 
but is still in the experimental state as far as boiler firing is concerned. 



Mixture of Pulverized 
Coal and Air 




^^^^^^^^^^^^^^^^^^^^^^^^^^^&^u 



Fig. 14. Blake Coal-dust Feeder. 



39. Blake Pulverizer and Feeder. — Fig. 14 gives a sectional view 
of the Blake apparatus, and is a typical example of a self-contained 



FUELS AND COMBUSTION 



63 



system. It comprises a multi-stage centrifugal pulverizer, coal hopper, 
conveyor and fan mounted on a single bedplate. Referring to the illus- 
tration, coal previously crushed to nut size is fed to the hopper from 
the bottom of which it is conveyed by an endless screw to the first 
stage of the pulverizer. The lumps are thrown out radially by centrif- 
ugal force, due to the rapidly revolving bats, and are reduced to a 
dust by percussion and attrition. The largest chamber contains a fan, 
the function of which is to draw the pulverized material successively 
from one chamber to another and finally deliver it to the discharge 
spouts. The air is drawn into the fuel chamber with the coal through 
passage A, and also through opening B around the shaft. After enter- 
ing the fan chamber, the mixture of coal dust and air receives an 
additional supply of air through opening C. The apparatus may be 
belt-driven or direct-connected and runs at about 1200 to 1600 r.p.m., 
requiring 8 to 12 horse power for its operation. Experience has demon- 
strated that as much as 14 per cent of moisture in the coal has little 
effect on the pulverization and burning. 

Fig. 11 shows a successful commercial application of the system to 
a 300-horse-power boiler. Table 16 gives the results of an evapora- 
tion test of this installation. 




Fig. 15. Triumph Coal-dust Feeder. 



40. Triumph Apparatus. — Fig. 15 illustrates the Triumph coal-dust 
feeder as designed by the C. O. Bartlett & Snow Company, Cleveland, 
Ohio. 

The coal is fed from storage bin to hopper A and feed worm B. The 
latter forces it down spout F directly to delivery tube D, where it is 
caught by the air draft and fed into the furnace. 

The amount of feed depends upon the speed of the feed worm, which 
is driven by a friction disk I against the flange plate H. This disk 
is moved in or out by handle, so as to get any speed desired. The 
air is furnished by fan C, the amount being controlled by valve E. 



64 STEAM POWER PLANT ENGINEERING 

41. Efficiency of Powdered-coal Furnaces. — A comparison of a num- 
ber of tests of hand-fired and powdered-coal furnaces with different types 
of feeders shows a decided gain in efficiency of the powdered coal over 
the hand-fired where the fuel is of a low grade. The gain becomes less 
marked with fuel of fair quality and disappears entirely with good fuel 
and properly manipulated automatic stokers. A test made by G. H. 
Barrus on a 250-horse-power B. & W. boiler at the General Electric 
Works in connection with a coal-dust feeder manufactured by the 
Phcenix Investment Company of New York gave a boiler and furnace 
efficiency of 75.3 per cent. A test of a 135-horse-power return tubular 
boiler with this same stoker gave a combined efficiency of boiler and 
furnace of 80 per cent. These figures, however, have been equaled and 
even exceeded in special hand-fired automatic-stoker tests, and only a 
comparative test of the two systems under similar conditions will show 
their respective efficiencies. 

Table 16 gives the results of a test of a 300-horse-power B. & W. 
boiler equipped with the Blake system. This plant represents the best 
practice in powdered-coal burning at this date (Jan., 1912). 

TABLE 16. 

TEST WITH PULVERIZED COAL AT THE HENRY PHIPPS POWER PLANT. 

Duration of test, hours •. . 6 

Total weight of coal fired, pounds 5,160 

Total weight of water, pounds 56,160 

Average temperature feed water, degrees Fahrenheit 186 , 

Average steam pressure, pounds per square inch 162.3 

Factor of evaporation 1.084 

Water evaporated per pound of coal 10.88 

Water evaporated per pound from and at 212 degrees, pounds . . 11 .725 

Boiler efficiency (coal containing 14,350 B.t.u.), per cent .... 78.93 

Horse power of boiler 294.6 

Builder's rating 300 

Temperature of escaping gases, degrees Fahrenheit 386 

Cost of coal, 2.50 tons at $1,315 per ton $3,392 

Cost of coal per pound 0.0006575 

Pounds of coal per boiler horse power per hour 2.92 

Cost of coal per boiler horse power, cents 0.192 

42. Rate of Combustion with Powdered Fuel. — In forcing large quan- 
tities of dust into the furnace the velocity imparted to the particles may 
be so great as to carry them beyond the zone of combustion before 
oxidation is complete, with the result that the flues and the back of the 
furnace become covered with unconsumed carbon. So much depends 
upon the depth of the furnace and the arrangement of the regenerative 
surface that no specific figures can be given as to the maximum rate of 



FUELS AND COMBUSTION 65 

combustion that can be efficiently effected. At ordinary rates of com- 
bustion the small particles of fuel are completely oxidized while in the 
combustion chamber and there is total absence of smoke. The use of 
anthracite coal is practically excluded from this type of stoker unless 
mixed with coal high in volatile matter. This is due to the fact that 
fixed carbon burns more slowly than the hydrocarbon gases and the 
temperature of ignition is higher; hence the most gentle draft will carry 
away the particles before they are completely consumed. With fuels 
high in volatile matter the hydrocarbons are distilled at a compara- 
tively low temperature, forming an inflammable gas which burns rapidly 
with the fixed carbon. A mixture of 30-per-cent bituminous and 70- 
per-cent anthracite has been successfully burned in the powdered form. 

43. Draft for Powdered Fuel. — A study of a number of tests of 
boilers burning powdered coal shows that the necessary draft is very 
low and ranges from 0.05 to 0.2 inch of water and averages not far from 
0.1 inch. 

44. Storing Powdered Fuel. — Most cities limit the storage of pow- 
dered coal to such a small quantity as to prohibit the use of fuel feeders 
of the " dust-feed" type in plants of any size not provided with a pul- 
verizing and crushing system. Coal dust mixed with air is often 
claimed to be of an explosive nature and many accidents are reported 
to have resulted from this cause. Many engineers, however, refute 
this on the basis of experiments which show that explosion can only 
occur at temperatures high enough to drive off the volatile gases. 

Explosibility of Coal Dust: Engr., Jan. 27, 1911; Mar. 31, 1911. Bureau of 
Mines, Bulletin No. 20, 1911. Fuel, Jan. 12, 1909, p. 294. 

45. Depreciation of Powdered-coal Furnaces. — To withstand the in- 
tense heat of combustion, brickwork of the highest quality is essential, 
since common fire brick are soon reduced to a liquid slag. A good 
quality of firebrick will withstand the heat for several months without 
renewals provided the furnace is properly enclosed, otherwise the strain 
of expansion and contraction due to alternate heating and cooling will 
crack the brick. Excellent results have been obtained from the use of 
bricks composed chiefly of the refuse from a carborundum slag, but the 
high cost has prevented their general use. 

46. Cost of Pulverizing Coal. — In stokers of the " Blake Pulverizer" 
type in which the grinding and feeding are carried on simultaneously 
in a self-contained apparatus, the power consumed varies from 2 to 
10 per cent of the total power developed, depending upon the nature 
of the fuel, the load factor, the efficiency of the driving mechan- 
ism, and the degree of fineness of the powdered fuel; 5 per cent is a 



66 STEAM POWER PLANT ENGINEERING 

fair average. The best results are obtained when 95 per cent of the 
dust will pass a 100 mesh and 75 per cent a 200 mesh, though satis- 
factory results have been obtained with as low as 40 mesh. Powdered 
coal in the open market ranges from 50 cents to 75 cents a ton above 
the price of the same coal in the form of screenings. 

Firing Boilers with Pulverized Coal: Power, Feb. 14, 1911. 

Pulverized Fuel: Eng. Mag., Jan., 1908; Jour. W. S. Engrs., Feb., 1904. 

Use of Pulverized Coal Under Steam Boilers: Power, Apr., 1904, March, 1904; 
Eng. and Min. Jour., Dec. 16, 1905; Col. Guard., Feb. 16, 1912. 

Tests of Pulverized Fuel: Engr. U. S., Apr. 1, 1904; Power, May, 1904, Feb. 14, 
1911. 

Types of Coal Dust Burners: Engr. U. S., Apr. 1, 1904; Jan. 1, 1903; Power, 
Mar., 1904. 

Burning Low Grade Coal Dust: Power, Sept. 12, 1911, p. 393. 

47. Fuel Oil. — The recent development of oil wells in the Western 
and Gulf States, with the consequent enormous increase in production, 
has given a marked impulse to the use of crude oil for fuel purposes in 
steam power plants. Where economic and commercial conditions per- 
mit, it is the most desirable substitute for coal. The total absence of 
smoke and ashes, prompt kindling and extinguishing of fires, extreme 
rate of combustion, and ease with which it can be handled and controlled 
are marked advantages in favor of fuel oil. The reduction in volume 
and weight over an equivalent quantity of coal for equal heating values 
and the increase in boiler efficiency are factors of no mean importance, 
particularly in connection with marine or locomotive work. In station- 
ary work the chief objections are the difficulty in securing ample storage 
capacity and the increased rate of insurance. An objection some- 
times raised against oil fuel is the increased depreciation of the setting, 
but in a well-designed setting this figure is only nominal and of second- 
ary importance. However, in spite of the many advantages presented 
in the use of fuel oil for power plant purposes, the comparatively limited 
supply prevents its adoption as a general fuel and limits its use to the 
plants most favorably located. 

48. Chemical and Physical Properties of Fuel Oil. — Crude oil, as 
pumped at the wells, consists principally of various combinations of 
hydrogen and carbon, together with small amounts of nitrogen, oxygen, 
sulphur, water in emulsion and silt. The nitrogen and oxygen may be 
classified with the moisture and silt as inert impurities. The moisture 
in oil fuel should not exceed 2 per cent, since it not only acts as an 
inert impurity, but must be converted into steam in the furnace and 
thus still further reduces the heat value per pound. The sulphur, 
though combustible, has a low calorific value and is otherwise unde- 
sirable. From Table 17 it will be seen that the physical properties 



FUELS AND COMBUSTION 



67 





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STEAM POWER PLANT ENGINEERING 



of oils from different localities in the United States differ widely, while 
the chemical constituents vary but slightly. For example, the oils 
given in the table differ greatly in volatility, specific gravity, and vis- 
cosity, but have approximately the same percentages of carbon and 
hydrogen. Taking hydrogen and carbon as the principal constituents 
it is found that oils rich in hydrogen are lighter in weight than those 
rich in carbon. Other things being equal, oils rich in hydrogen have a 
higher calorific value than those rich in carbon, but the heavier oils are 
usually the cheaper. The relation between heating value and specific 
gravity for anhydrous California oil is as follows: 

TABLE 18. 



APPROXIMATE RELATION BETWEEN THE HEATING VALUE AND SPECIFIC GRAVITY. 




(Professor Le Conte, University of California.) 




Degrees, 


Specific 


Weight per 


B.t.u. per 


B.t.u. per 


Degrees, 


Baume. 


Gravity. 


Barrel. 


Pound. 


Barrel. 


Baum6. 


10 


1.0000 


350.035 


18,280 


6,398,600 


10 


11 


0.9929 


347.55 


18,340 


6,374,100 


11 


12 


0.9859 


345.10 


18,400 


6,349,800 


12 


13 


0.9790 


342.68 


18,460 


6,325,900 


13 


14 


0.9722 


340.30 


18,520 


6,302,400 


14 


15 


0.9655 


337.96 


18,580 


6,279,300 


15 


16 


0.9589 


335.65 


18,640 


6,256,500 


16 


17 


0.9524 


333.37 


18,700 


6,234,000 


17 


18 


0.9459 


331.10 


18,760 


6,211,400 


18 


19 


0.9396 


328.89 


18,820 


6,189,700 


19 


20 


0.9333 


326.69 


18,880 


6,167,900 


20 • 


21 


0.9272 


324.55 


18,940 


6,147,000 


21 


22 


0.9211 


322.42 


19,000 


6,126,000 


22 


23 


0.9150 


320.28 


19,060 


6,104,500 


23 


24 


0.9091 


318.22 


19,120 


6,084,400 


24 


25 


0.9032 


316.15 


19,180 


6,063,800 


25 



The heat value may be closely approximated by means of the follow- 
ing formula (Jour. Am. Chem. Soc, Oct., 1908): 

B.T.U. = 18,650 + 40 (B - 10), 
in which B = degrees Baume. 

Oil that is to be transported or stored or used for fuel inside of build- 
ings should be of the "reduced" variety, from which the naphtha and 
higher illuminating products have been distilled. The gravities of 
such distillates vary from 20 to 25 degrees Baume, or close to 0.9 specific 
gravity, and their flash points range from 240 degrees F. to 270 degrees F. 
One barrel of crude oil contains 42 gallons and weighs from 310 to 
350 pounds, according to the specific gravity. Compared with coal, oil 
occupies about 50 per cent less space and is 35 per cent less in weight, 



FUELS AND COMBUSTION 



69 



for equal heat values. The comparative heat values of coal and oil are 
approximately as follows: 



B.T.U. per Pound 
of Coal. 


Pounds of Coal Equal 
to 1 Barrel of Oil. 


Barrels of Oil Equal 

to 1 Short Ton of 

Coal. 


10,000 


620 


3.23 


11,000 


564 


3.55 


12,000 


517 


3.87 


13,000 


477 


4.19 


14,000 


443 


4.52 


15,000 


413 


4.84 



49. Efficiency of Boilers with Fuel Oil. — A coal-burning boiler 
which utilizes 80 per cent of the heat value of the fuel is exceptional — ■ 
75 per cent represents very good practice, and 70 per cent a fair aver- 
age for good practice. The great majority of coal-burning boilers, 
however, operate at efficiencies less than 65 per cent. With oil fuel a 
boiler and furnace efficiency of 75 per cent is quite ordinary and 80 per 
cent not uncommon. This increase in efficiency is partly due to the 
fact that the oil is readily broken up and brought into immediate con- 
tact with the necessary air for combustion and loss due to excessive 
air dilution is correspondingly reduced. 

Table 19 gives the theoretical air requirements for different densities 

of fuel oils and Table 20 the air excess for various efficiencies. These 

tables were compiled by C. R. Weymouth (Trans. A.S.M.E., Vol. 30, 

p. 801). 

TABLE 19. 



POUNDS OF AIR PEE 


POUND OF OIL AND RATIO OF AIR SUPPLIED TO THAT 






CHEMICALLY REQUIRED. 








Light Oil, 


Med 


ium Oil, 


Heavy Oil, 




C, 84%; H 


, 13%; S, 0.8%; 


C, 85%; H 


, 12%; S, 0.8%; 


C, 86%; H 


, 11%; S, 0.8%; 


Per Cent C0 2 


N, 0.2%; 


, 1%; H 2 0, 1%. 


N, 0.2%; 'O 


1%; H 2 0, 1%. 


N, 0.2%; O 


, 1%; H 2 0, 1%. 


by Volume as 














Shown by 














Analysis of Dry 
Chimney Gases. 


Lbs. of 

Air per 

Lb. of Oil. 


Ratio of Air 
Supply to 


Lbs. of 

Air per 

Lb. of Oil. 


Ratio of Air 
Supply to 


Lbs. of 

Air per 

Lb. of Oil. 


Ratio of Air 
Supply to 




Chemical 
Requirements. 


Chemical 
Requirements. 


Chemical 
Requirements. 


4 


51.40 


3.607 


51.93 


3.704 


52.45 


3.803 


5 


41.31 


2.899 


41.71 


2.975 


42.12 


3.054 


6 


34.58 


2.427 


34.90 


2.490 


35.23 


2.554 


7 


29.77 


2.089 


30.04 


2.143 


30.31 


2.198 


8 


26.17 


1.836 


26.39 


1.883 


26.62 


1.930 


9 


23.37 


1.640 


23.56 


1.680 


23.75 


1.722 


10 


21.12 


1.482 


21.29 


1.518 


21.45 


1.555 


11 


19.83 


1.391 


19.43 


1.386 


19.58 


1.419 


12 


17.76 


1.246 


17.88 


1.276 


18.01 


1.306 


13 


16.46 


1.155 


16.57 


1.182 


16.69 


1.210 


14 


15.36 


1.078 


15.45 


1.102 


15.55 


1.127 


15 


14.39 


1.010 


14.48 


1.033 


14.57 


1.056 



70 



STEAM POWER PLANT ENGINEERING 



Table 21 gives the results of a series of tests made at the Redondo 
plant of the Pacific Light & Power Company, California, on a 604-horse- 
power B. & W. boiler equipped with Hammel furnaces and burners. 
The boiler was in regular service and under usual operating conditions. 

TABLE 20. 

BOILER EFFICIENCY FOR EXCESS AIR SUPPLY (OIL FUEL). 



Excess Air Supply, Per Cent. 



Assumed temperature of escaping gases, 
deg. F 

Corresponding ideal efficiency of boiler, 
per cent 

Possible saving in fuel due to reduction 
of air supply to 10 per cent excess, ex- 
pressed as per cent of oil actually 
burned under assumed conditions .... 



10 


50 


75 


100 


150 










Over 


400 


450 


475 


490 


500 
Under 


84.2 


80.27 


77.66 


75.22 


70.94 
Over 





4.67 


7.78 


10.68 


15.76 



200 



Over 
500 
Under 

67.09 



Over 
20.32 



50. Comparative Evaporative Economy of Oil and Coal. — In deter- 
mining the comparative economy of coal and oil, the fixed and operating 
charges must be considered in addition to the cost and efficiency of 
the fuel. From the market quotation on oil and coal and the com- 
parative heating values of each the actual cost per B.t.u. is readily 
obtained, and by combining this with the relative efficiencies from the 
furnace standpoint the net cost of the fuel is obtained. The fixed 
charges vary with the location and size of the plant and are approxi- 
mately the same per boiler horse power for a given location in both 
cases. The insurance rates may be greater with the oil fuel and the 
depreciation of the boiler setting may be somewhat larger, but in a well- 
constructed furnace the latter item should be the same in both instances 
for average rates of combustion. The operating charges are decidedly 
in favor of the oil fuel, since no ash handling is necessary. Oil fuel is 
readily fed to the furnace, and the cost of attendance may be materially 
less than with coal firing, and one man may safely control from eight to 
ten boilers. Table 148, Chapter XVII, gives data relative to the cost 
of producing electrical power in connection with oil-fired steam plants. 

51. Oil Burners. — The function of the burner is to atomize the oil 
to as nearly a gaseous state as possible. 

Classification of a few well-known burners 
Mechanical Spray: 

Korting. 
Vapor or Carburettor: 

Durr. 

Harvey. 



FUELS AND COMBUSTION 



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72 STEAM POWER PLANT ENGINEERING 

Spray Burners: 
Outside Mixers. 

(a) Peabody. 

(b) Warren. 

Inside Mixers. 

(a) Hammel. 

(b) Kirkwood. 

(c) Branch. 

(d) Williams. 

Oil burners for burning liquid fuel may be divided into three general 
classes : 

1 . Mechanical spray, in which the oil, previously heated to a temper- 
ature of about 150 degrees F., is forced under pressure through nozzles 
so designed as to break it up into a fine spray. The Korting Liquid 
Fuel Burner, Fig. 16, is an example of this type. In this design a 




Fig. 16. Korting Fuel-oil Burner. 

central spindle, spirally grooved, imparts a rotary motion to the oil 
and causes it to fly into a spray by centrifugal force on issuing from 
the nozzle. The particles of oil are burned in the furnace when they 
come in contact with the necessary air to effect combustion. This 
type of burner is little used in this country in connection with power- 
plant work, but is meeting with much success on the continent. 

2. Vapor burners, or carburettors, in which the oil is volatilized in 
a heater or chamber and then admitted to the furnace, are seldom used 
except in connection with refined oils, as the residuals from crude oil 
are vaporized only at a high temperature. The Durr and Harvey 
gasifiers are the best known of this type. 

3. Spray burners are by far the most common in use. In this type 
the oil is held in suspension and forced into the furnace by means of a 
jet of steam or compressed air. Spray burners are designed either as 
outside mixers, in which the oil and atomizing medium meet outside the 
apparatus, or inside mixers, in which the oil and atomizing medium 
mingle inside the apparatus. 






FUELS AND COMBUSTION 



73 



The Peabody burner, Fig. 17, illustrates the principles of the "outside- 
mixer" type of apparatus. In this type the oil flows through a thin 
slit and falls upon a jet of steam which atomizes it and forces it into the 



rm 




EH 



Fig. 17. Peabody Fuel-oil Burner. 

furnace in a fan-shaped spray. A feature of this apparatus is its 
simplicity of construction. 

Fig. 18 illustrates the Hammel burner as used at the power house of 
the Pacific Light and Power Company, Los Angeles, Cal. Oil enters the 
burner under pressure and flows through opening D to the mouth of 
the burner, where it is atomized by the steam jets issuing from slots G, 




o 


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Fig. 18. Hammel Fuel-oil Burner. 

H, and J. The oil is preheated to facilitate its flow through the supply 
system. Plates K-K are removable and are easily replaced when worn 
out or burned. The Hammel burner belongs to the " inside mixers." 
A few well-known types of "inside mixers" are illustrated in Figs. 18 
to 20. The operation is practically the same in all of them and they 
differ only in mechanical details. 



74 



STEAM POWER PLANT ENGINEERING 




Fig. 20. Kirkwood Fuel-oil Burner. 




Fig. 21. Williams Fuel-oil Burner. 



FUELS AND COMBUSTION 



75 



The Williams burner, Fig. 21, differs somewhat from the others in 
that the air supply passes through the burner and mingles with the oil 
and steam before entering the furnace. 

The simplest and most reliable burners are of the Hammel type and 
are much in evidence in the Pacific States. 

52. Furnaces for Burning Oil Fuel. — The efficient combustion of 
oil fuel depends more upon the proportions of the furnace than upon 
the type of burner, provided, of course, the latter is of modern design. 
While it is desirable to have incandescent brickwork around the flame 
it is impossible to do so in many cases and a satisfactory compromise 




Fig. 22. Warren Fuel-oil Burner. 



is effected by using a flat flame burning close to a white-hot floor through 
which air is steadily flowing. A good burner will maintain a suspended 
flame clear and smokeless in a cold furnace. The path of the flame 
in the furnace must be such as to insure uniform distribution of heat 
over the boiler heat-absorbing surfaces without direct flame impinge- 
ment. Under ordinary firing the flame should not extend into the 
tubes. The first pass of the boiler should be located directly over the 
furnace in order that the heating surface may absorb the radiant energy 
from the incandescent fire brick. Fire-brick arches and target walls 



76 



STEAM POWER PLANT ENGINEERING 



are not to be recommended on account of the localization of heat re- 
sulting in burning out the tubes or bagging the shell and on account of 
the limited overload capacity. 

Fig. 23 shows the general details of a Hammel oil-burning furnace 
illustrating current practice on the Pacific coast. The burner tip is 
housed in a slot located in the back of an arched recess in the bridge 




Fig. 23. Furnace for Burning Fuel Oil, Rear Feed (Hammel). 

wall and the flame is projected forward toward the front of the furnace. 
The furnace floor is carried on pieces of old two-inch pipe or on old 
rails and is solid except for narrow air slots through the deck and in 
front of each arch. Each burner with its accompanying recess has a 
separate air tunnel from the boiler front; these tunnels do not commu- 



FUELS AND COMBUSTION 



77 



nicate with each other under the furnace floor and by closing the ash- 
pit door any tunnel can be sealed up while the others are supplying air 
to their particular burners. The Hammel furnace is a modification of 




Fig. 24. — Peabody Fuel Oil Furnace. 




Fig. 25. — Modern Furnace for Burning Fuel Oil, Front Feed. 



the well-known Peabody furnace, a section through which is shown in 
Fig. 24. 

Fig. 25 gives the general details of a modern oil-burning furnace, 
with front feed, as applied to a horizontal return tubular boiler. 



78 



STEAM POWER PLANT ENGINEERING 



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FUELS AND COMBUSTION 79 

53. Atomization of Oil. — For efficient combustion the oil should be 
injected into the furnace in the form of a spray. Three systems of 
atomization are in use in stationary practice, namely, mechanical, air, 
and steam. Of these, by far the greater number of installations in the 
United States are of the last order. 

The mechanical or Korting system is not much in evidence in this 
country, but it is used extensively in Europe. The operation of this 
system is described in paragraph 51. The makers state that to oper- 
ate the pumps and supply the heat to the oil takes from f to 1 per cent 
of the steam evaporated. Mechanical atomization presents many pos- 
sibilities and it is not unlikely that future development may lie along 
this path. 

In air atomization the air is used at pressures from 1J to 60 pounds 
per square inch, depending upon the type of burner. From Table 22 
it will be seen that the total steam used to compress the air varies from 
1.06 to 7.45 per cent of the total generated. For air atomization and 
with air pressures of from 20 to 30 pounds per square inch, J. H. Hoppes, 
(Jour. A.S.M.E., Aug., 1911, p. 902), states that from 6 to 10 cubic 
feet of air per minute per pound of oil burned will be required. Com- 
pressed air offers no opportunity for fuel saving over the use of steam 
direct in cases where steam is available. In certain industrial oper- 
ations where high temperatures are essential the use of air is preferred. 
When it is necessary to use high-pressure air the economy decreases 
with the increase in pressure, since the cost of each cubic foot of com- 
pressed air increases rapidly with the pressure, but its ability to atomize 
the oil does not increase proportionately. 

Steam is the most commonly used medium for atomizing the oil, 
since its use obviates complication and risk of interrupted service. 
The amount of steam required to atomize the oil varies from 0.3 to 
0.7 pounds per pound of oil, with an average of about 0.5 pounds. The 
steam consumption is generally stated in per cent of the total steam 
generated, but the results are misleading since the percentage factor 
depends largely upon the efficiency of the boiler. Table 22 gives the 
results of a number of tests of different types of burners with air and 
steam as an atomizing medium. 

54. Oil-feeding Systems. — Fig. 26 gives a diagrammatic arrange- 
ment of the piping commonly employed in feeding oil fuel to the burners. 
Steam-actuated oil pumps, installed in duplicate, draw the fuel from 
the supply tank and deliver it under pressure to the burners. The 
piping is cross-connected so that repairs can be made without inter- 
rupting the service. The oil is heated from the pump exhaust before 
it is supplied to the burners. This should not be carried beyond the 



80 



STEAM POWER PLANT ENGINEERING 




a 
1 
ft 

•3 



bO 



FUELS AND COMBUSTION 81 

flash point of the oil used or there will be danger from carbon deposits 
in the supply pipe. A strainer is placed in the suction line between 
the storage tank and the oil-pressure pump to minimize clogging of 
the burner. In some instances strainers are also placed in the supply- 
pipe between the heater and burner. The relief valve between the 
pumps and burners is set at a definite maximum oil pressure so as to 
prevent excessive pressure. The oil meter is for the purpose of check- 
ing the storage tank indicator. All oil piping is installed so that it 
can be drained back to the storage tank by gravity in case of neces- 
sity. In many large plants the strainers, meters, heaters and piping 
are installed in duplicate. Arrangements are usually made for the oil 
to be delivered at constant pressure. The supply of steam to the 
burner is controlled by regulating the pressure in a separate main 
common to all burners, the pressure in the main bearing a certain pre- 
determined relation to the pressure in the oil mains. In most installa- 
tions the suppfy of steam and oil at the burner is regulated by hand to 
meet the requirements of the individual burners. At the Redondo 
plant of the Pacific Light and Power Company, Redondo, CaL, the 
supply of oil and steam to all burners and the supply of air for com- 
bustion to any number of boilers are automatically controlled from a 
central point. For a description of this system see Trans. A.S.M.E., 
Vol. 30, p. 808. 

Low-pressure systems are ordinarily operated under standpipe 
pressures as in Fig. 27, which illustrates the arrangement of apparatus 
as advocated by the International Gas and Fuel Company. A steam 
pump B draws the oil from the buried tank through pipe Z and delivers 
it to the standpipe E. Thence it flows through pipe I to the burners 
under a head of about 10 feet. The pump runs constantly, the surplus 
oil flowing back to the tank through the pipe T. The oil is heated by 
the exhaust pipe Z' . The oil pump is provided with a device D having 
a piston connected by a chain with a cock S, which automatically opens 
when the boiler is not under steam pressure, so that the standpipe will 
be emptied, the oil flowing to the storage tank. 

Fig. 28 illustrates the Hydraulic Oil Storage Company's system of 
storing oil and delivering it to the burners. The oil reservoirs are 
placed below grade, as indicated, to minimize fire risk. The operation 
is as follows: Water enters the " float box" and flows through a "three- 
way cock" to the bottom of the reservoir until all of the oil and water- 
pipes are filled up to the level of the float box, when the float auto- 
matically cuts off the supply. This flooding of the entire system drives 
out all of the air. The three-way cock is then turned to " discharge" 
and part of the water flows to the sewer. The tank car or wagon is 



82 



STEAM POWER PLANT ENGINEERING 



next attached to the "oil inlet" and the oil flows into the tank and dis- 
places the water until the level of the " filler float" is reached, when the 
supply is automatically cut off. The inlet is so placed that the head 
of oil in the tank car is sufficiently great to overcome the opposing 
head of water. The three-way valve is next turned to the first position 
and the head of water forces the oil to the burners. After the oil has 
been withdrawn from the storage tank the water can only rise to the 
level of the water in the float box and therefore cannot be fed to the 
furnace. The small steam pipe admits steam into the tank and heats 
the oil, thereby making it flow more freely. 




Fig. 27. International Gas and Fuel Company's Fuel Oil System. 

55. Oil Transportation and Storage. — Fuel oil is delivered in bulk, 
either in tank cars, barges or steamships, or by pipe lines, depending 
upon the location of the plant. It must be stored in accordance with 
underwriters' requirements and community ordinances. In outlying 
districts the storage reservoirs may be placed above the ground, but 
in cities they must be located underground. In small plants surface 
tanks are sometimes constructed of wood but in the majority of in- 
stallations both surface and underground tanks are constructed of 
steel plate, plain concrete or reinforced concrete. In large plants the 
greater portion of the oil is stored in surface tanks some distance away 



FUELS AND COMBUSTION 



83 



from the plant. From these main reservoirs the oil is pumped into 
auxiliary tanks or vaults outside the building, but beneath the boiler- 
room level. In no case is the oil permitted to gravitate to the burners 
but must reach them by means of pumps. Storage tanks should be 
fitted with vent pipes; return and overflow pipes; indicators showing 
the depth of oil; steam coils for preliminary heating in cold weather, 
or, for thick, viscous oils, steam pipes for smothering the flames in case 
of fire, and suitable manholes for cleaning out purposes. To conform 
with underwriters' requirements the tops of the tanks should be placed 



Oil Inlets G 



Filler Float 



, L .•'"Oil Supply Pipe 



DC 



f 



I Deflecto 

\ 

| Oil Res 

W//////W/////////77 / 



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srvoir 



7 7777M 







flj^vSiv^ 



Steam +? 



?, Deflector 

I 



Fig. 28. Hydraulic Oil Storage Company's Fuel Oil System. 



below the level of the lowest pipe used in connection with the apparatus. 
Steel tanks, including foundations, roof and setting, cost approximately 
$1.00 per barrel, while concrete vaults and forms cost approximately 
fifty cents per cubic foot of concrete exclusive of excavation. 

56. The Purchase of Fuel Oil. — The following extracts from Bulletin 
No. 3, 1911, Bureau of Mines (" Specifications for the Purchase of Fuel 
Oil for the Government, with Directions for Sampling Oil and Natural 
Gas"), though primarily intended for the guidance of Government 
officials, may be of service to engineers: 



84 STEAM POWER PLANT ENGINEERING 

1. In determining the award of a contract, consideration will be 
given to the quality of the fuel offered by the bidders, as well as the 
price, and should it appear to be the best interest of the Government 
to award a contract at a higher price than that named in the lowest 
bid or bids received, the contract will be so awarded. 

2. Fuel oil should be either a natural homogeneous oil or a homo- 
geneous residue from a natural oil ; if the latter, all constituents having 
a low flash point should have been removed by distillation; it should 
not be composed of a light oil and a heavy residue mixed in such pro- 
portions as to give the density desired. 

3. It should not have been distilled at a temperature high enough 
to burn it nor at a temperature so high that flecks of carbonaceous 
matter began to separate. 

4. It should not flash below 60 degrees C. (140 degrees F.) in a closed 
Abel-Pensky or Pensky-Martens tester. 

5. Its specific gravity should range from 0.85 to 0.96 at 15 degrees C. 
(59 degrees F.); the oil should be rejected if its specific gravity is above 
0.97 at that temperature. 

6. It should be mobile, free from solid or semi-solid bodies, and 
should flow readily at ordinary atmospheric temperatures and under 
a head of 1 foot of oil, through a 4-inch pipe 10 feet in length. 

7. It should not congeal nor become too sluggish to flow at degree C. 
(32 degrees F.). 

8. It should have a calorific value of not less than 10,000 calories 
per gram (18,000 B.t.u. per pound); 10,250 calories to be the standard. 
A bonus is to be paid or a penalty deducted according to the method 
stated under section 21, as the fuel oil delivered is above or below this 
standard. 

9. It should be rejected if it contains more than 2 per cent water. 

10. It should be rejected if it contains more than 1 per cent sulphur. 

11. It should not contain more than a trace of sand, clay or dirt. 

12. Each bidder must submit an accurate statement regarding the 
fuel oil he proposes to furnish. This statement should show: 

(a) The commercial name of the oil. 

(b) The name or designation of the field from which the oil is obtained. 

(c) Whether the oil is a crude oil, a refinery residue, or a distillate. 

(d) The name and location of the refinery, if the oil has been re- 
fined at all. 

For sampling, analysis, etc., consult complete bulletin. 

Analyses of California Petroleums: Bulletin No. 19, U. S. Bureau of Mines, 1912. 
Atomization: Jour. A.S.M.E., Aug. 11, 1911, p. 883, 902; Jl. El. Power and Gas, 
Dec. 23, 1911. 



FUELS AND COMBUSTION 



85 



Burners: Jl. El. Power and Gas, Dec. 23, 1911, Apr. 1, 1911; Engng.,' Feb. 16, 
1912. 

Comparative Evaporative Value of Coal and Oil: Jl. El. Power and Gas, March 18, 
1911; Jour. A.S.M.E., Aug. 11, 1911, p. 872. 

Draft Requirements for Burning Oil Fuel: Jour. A.S.M.E., Aug., 1911; Oct., 1912. 

Economy Tests with Oil Fuel: Trans. A.S.M.E., 30-1908, p. 775; Jl. A.S.M.E., 
Aug. 11, 1911, p. 940. 

Furnqces for Burning Oil Fuel: Jl. El. Power and Gas, Dec. 30, 1911, Apr. 8, 
1911; Jour. A.S.M.E., Aug., 1911, p. 879; Ir. Td. Review, June 3, 1908; Power, 
June 16, 1908. 

Oil for Steam Boilers: Jour. A.S.M.E., Aug., 1911, p. 931; Jl. El. Power and 
Gas, Dec. 16, 1911; Power, Aug., 1908, p. 943; Jan. 23, 1908, p. 980; Bulletin 
No. 131, Louisiana State University. 

Precautions with Oil Fuel: Eng. and Min. Jour., Apr. 1, 1911, p. 653. 

Purchase of Fuel Oil for the Government: Bulletin No. 3, Bureau of Mines, 1911. 

Regulation of Oil Supply to Burners: Trans. A.S.M.E., 30-1908, p. 804. 

Storage and Transportation: Jl. El. Power and Gas, Dec. 16, 1911, p. 564; Eng. 
News, Sept. 25, 1902, p. 232; Power, July 16, 1908. 

Unnecessary Losses in Firing Fuel Oil: Trans. A.S.M.E., 30-1908, p. 797. 

57. Gaseous Fuels. — These fuels offer all of the advantages of liquid 
fuels and but few of the disadvantages. The gases most commonly met 
with in connection with steam power plants are outlined in Table 23. 
The artificial gases for steam purposes are prohibitive in cost in most 
cases, and even in blast-furnace installations, where the gases are waste 
products, the gas engine has virtually supplanted the steam engine for 
power purposes. In the immediate locality of natural-gas wells gas- 
fired furnaces may prove to be more economical than coal furnaces, but 
the limited supply limits its use as a general fuel. From the market 
quotations on coal and gas and the comparative heating value of each 
the actual cost per B.t.u. is readily obtained, and by combining this 
with the relative efficiencies from the furnace standpoint the net cost 
of the fuel is obtained. The following, table, based upon the assump- 
tion that one cubic foot of natural gas under standard conditions has a 
heating value of 1000 B.t.u., will enable an approximate comparison to 
be made : 



B.T.U. per Pound of 


Pounds of Coal Equal 


No. of 1,000 Cu. Ft. 


Coal 


to 1,000 Cu. Ft. 


of Gas Equal to One 




of Gas. 


Short Ton of Coal. 


10,000 


100 


20 


11,000 


91 


22 


12,000 


83 


24 


13,000 


77 


26 


14,000 


71 


28 


15,000 


67 


30 



86 



STEAM POWER PLANT ENGINEERING 



In burning natural gas under a boiler the furnace requirements are 
practically the same as for liquid fuel. The burners, of course, will 
differ in design, since atomization is unnecessary. The majority of 
patented gas burners are operated on the same principle as the com- 
mon gas-stove burner. A crude form often used consists of a J-inch 
gas pipe placed within a 2J-inch pipe which is bricked in the fire-door 
opening. A properly installed gas-fired furnace should be capable of 
converting 72 per cent of the heat value of the fuel into steam, cor- 
responding to approximately 80 per cent of the lower heating value. 

Fig. 29 shows a section through a small experimental boiler de- 
signed by Prof. Wm. A. Bone, University of Leeds, England, which in- 
volves the principle of so-called " surface combustion," and for which 




Fig. 29. Experimental Boiler Involving the Principles of " Surface Combustion." 

extravagant claims have been made as regards efficiency and capacity. 
It consists essentially of a plain tubular boiler, having ten tubes, 3 
inches in internal diameter. Each of these is bushed with a short tube, 
E, of fire clay and is filled for the rest of its length with finely broken 
refractory material. Mixing chambers of special design are at- 
tached to the front plate of the boiler as indicated. The mixture fed 
into the boiler tubes from these mixing chambers consists of the com- 
bustible gas with a proportion of air very slightly in excess of that 
theoretically required for combustion. The mixture is injected or 
drawn in through the orifice in the fire-clay plug. The gas burns with- 
out flame in the front end of the tube, the incandescent mass being in 
direct contact with the heating surface. The combustion of the mix- 
ture in contact with the incandescent material is completed before it 
has traversed a length of 6 inches from the point of entry of the tube. 



FUELS AND COMBUSTION 



87 






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88 STEAM POWER PLANT ENGINEERING 

Although the core of the material at this part of the tube is incan- 
descent the heat transference is so rapid that the walls of the tubes 
are considerably below red heat. The evaporation in regular working 
order is over 20 pounds per square foot of heating surface and this can 
be increased 50 per cent with a reduction in efficiency of only 5 or 6 
per cent. The figures given by Prof. Bone for the boiler and econo- 
mizer are as follows: 

Date, Dec. 8, 1910. 

Pressure of mixture entering boiler tubes, inches of water 17.3 

Pressure of products entering economizer, inches of water 2.0 

Steam pressure, pounds per square inch gauge 100.0 

Temperature of steam in boiler, degrees F 334.0 

Temperature of gases leaving boiler, degrees F 446.0 

Temperature of gases leaving economizer, degrees F 203.0 

Temperature of water entering economizer, degrees F 41.9 

Temperature of water leaving economizer, degrees F 136.4 

Evaporation per square foot of heating surface per hour, pounds 21.6 

Gas consumption, cubic feet per hour, at 32 degrees F. and 14.7 pounds 

per square inch 996.0 

B.t.u. per standard cubic foot (lower heat value) 562.0 

Water evaporated per hour from and at 212 degrees F., pounds 550.0 

Efficiency of boiler and economizer (on basis of low heat value), per 

cent 94.3 

For further details of Prof. Bone's experiment see American Gas Light Journal, 
Dec. 4, 1911; Engineering, April 14, 1911; Engineer (London), April 14, 1911. 
See also editorial, Industrial Engineering, Jan., 1912, p. 59. 



CHAPTER III. 

BOILERS. 

58. As affecting fuel economy the boiler equipment is by far the 
most important part of the power plant and involves the largest share 
of the operating expenses. It matters little how elaborate, modern, or 
well designed it may be, skill, good judgment, and continued vigilance 
are required on the part of the operator to secure the best efficiency. 

Of the various types and grades of boilers on the market experience 
shows that most of them are capable of practically the same evapora- 
tion per pound of coal, provided they are designed with the same pro- 
portions of heating and grate surface and are operated under similar 
conditions. They differ, however, with respect to space occupied, 
weight, capacity, first cost, and adaptability to particular conditions 
of operation and location. 

59. Classification. — As to design and construction there is an almost 
endless variety of boilers and furnaces, classified as internally and 
externally fired; water tube and fire tube; through tube and return tubular; 
horizontal and vertical. 

The internally fired type includes the vertical tubular, locomotive, 
Scotch-marine, and practically all flue boilers. The externally fired 
includes the plain cylinder, the through tubular, return tubular, and 
nearly all stationary water-tube boilers. 

60. Vertical Tubular Boilers. — Vertical tubular boilers, Figs. 1 and 
30, are commonly used where small power, compactness, low first cost, 
and sometimes portability are the chief requirements, though they 
are not necessarily restricted to small sizes. The tubes are sometimes 
arranged so that the spaces between them radiate from a hand hole on 
one side so that a scraper may readily be inserted to clean the top of 
the furnace plate. The hand hole in the water leg permits removal of 
the scale. It is convenient to place a chain in the bottom of the water 
leg, which can be worked around through the hand hole for the pur- 
pose of loosening up the scale deposit. The distance between the 
furnace crown and top of the grate is never less than 24 inches even in 
the smallest boiler and should be as great as possible to insure good 
combustion. Two styles of vertical boilers are in common use, the 
ordinary vertical type, Fig. 1, and the submerged type, Fig. 30. In 

89 



90 



STEAM POWER PLANT ENGINEERING 



the former the upper tube sheet and part of the tubes are above the 
water line, and while this feature may tend to superheat the steam to 
a slight extent, the difficulty from unequal expansion and liability to 
overheating is of sufficient moment to justify the use of the submerged 
type, particularly where the boiler is likely to be forced above its rated 

capacity. The advantages 
of this type of boiler are: 
(1) compactness and port- 
ability; (2) requires no 
setting beyond a light 
foundation; (3) is a rapid 
steamer, and (4) is low in 
first cost. The disadvan- 
tages are: (1) inaccessibility 
for thorough inspection and 
cleaning; (2) small steam 
space, which results in ex- 
cessive priming at heavy 
loads; (3) poor economy 
except at light loads, as 
the products of combustion 
escape at a high temperature 
on account of the shortness 
of the tubes; (4) smokeless 
combustion practically im- 
possible with bituminous 
coals; (o) the small water 
capacity results in rapidly 
fluctuating steam pressures 
with varying demands for 
steam. 

Although vertical fire- 
tube boilers are usually of 
very small size, being seldom 
constructed in sizes over 60 
horse power, an exception is 
found in the Manning boiler, Fig*. 31, which is constructed in sizes as 
large as 250 horse power. Many of the disadvantages found in the 
smaller types are obviated in the Manning boilers, which, as far as 
safety and efficiency are concerned, rank with any of the other first- 
class types. They differ from the boiler described above mainly in 
having the lower or furnace portion of much greater diameter than the 




Fig. 30. 



Blow 
Off 



Vertical Tubular Boiler with Submerged 
Tube Sheet. 



BOILERS 



91 



wbw^ ^TOWtob 



upper part which encircles the tubes. This permits a proper propor- 
tion of grate, which is not obtainable in boilers like Figs. 1 and 30. The 

double-flanged head connecting the 
upper and lower shells allows suf- 
ficient flexibility between the top 
and bottom tube sheets to provide 
for unequal expansion of tubes and 
shell. The ash pit is built of brick 
and the water leg does not extend 
below the grate level, thus doing 
away with dead-water space. Where 
overhead room permits and ground 
space is expensive, this boiler offers 
the advantage of taking up a small 
floor space as compared with hori- 
zontal types. 

61. Fire-box Boilers. — Although 
vertical fire-tube boilers may be 
classed as fire-box boilers, yet the 
term "fire box" is usually associated 
with the locomotive types, whether 
used for traction or stationary pur- 
poses. The usual form of fire-box 
boiler as applied to stationary work 





Manning Vertical Fire-tube Boiler. 



is illustrated in Fig. 32. The shell is prolonged beyond the front tube 
sheet to form a smoke box. The front ends of the tubes lead into the 
smoke box and the rear ends into the furnace or fire box. The fire box 



92 



STEAM POWER PLANT ENGINEERING 



is ordinarily of rectangular cross section, and is secured against collapse 
by stay bolts and other forms of stays. In Fig. 32 the smoke box is of 
cylindrical cross section and hence requires no staying except at the 
flat surface. Fire-box boilers are used a great deal in small heating 
plants where space limitation precludes other types. Their steam 
capacity gives them an advantage over the vertical tubular form. 
Being internally fired no brick setting is required. They are usually 



Safety Valve 




Fire Door 



Ash Door 



Fig. 32. Typical Fire-box Boiler — Stationary Type. 

of cheap construction, designed for low pressure, and seldom made in 
sizes over 75 horse power. Unless carefully designed and constructed 
high steam pressures are apt to cause leakage because of unequal expan- 
sion of boiler shell, tubes, and fire box. Portable fire-box boilers with 
return tubes are made in sizes as large as 150 horse power and for 
pressures as high as 150 pounds per square inch, but being more costly 
than some of the other types of boilers of equal capacity are used only 
where portability is an essential requirement. 

62. Fitzgibbons Boiler. — Fig. 33 shows a section through a Fitz- 
gibbons boiler and setting illustrating a combination of the vertical 
tubular and the locomotive fire-box type. This combination provides 
a large and efficient combustion chamber with economy of floor space. 
The horizontal tube sheets are completely submerged and the arrange- 
ment of the heating surface effects an exceedingly rapid water circula- 
tion. Fitzgibbons boilers are made in various sizes ranging from 10 
to 350 horse power and are finding favor with engineers for small 
installations. They are much in evidence in public buildings and 
institutions. 



BOILERS 



93 



63. Scotch-marine Boiler. — Where an internally fired boiler is 
desired for large powers the Scotch-marine type is finding much favor 
with engineers. A number of the tall office buildings in Chicago are 
equipped with boilers of this class which are giving good results. They 
require little overhead room, no brick setting, and are excellent steamers. 
The Continental boiler, Fig. 34, is one of the best known of this type. 
The boiler is self-contained and requires no brick setting, the only fire 
brick used being those that form the bridge wall, baffle ring, and the 
layer at the back of the combustion chamber. The furnace and tubes 
are entirely surrounded by water, so that all fire surfaces, excepting the 




Fig. 33. 100-horse-power Fitz gibbons Boiler. 



rear of the combustion chamber, are water cooled. The furnace is 
corrugated for its whole length. These corrugations, in addition to 
giving greater strength to the furnace, act as a series of expansion joints, 
taking up the strains due to unequal expansion of furnace and shell. 
Practically all types of mechanical stokers and grates are applicable to 
these boilers. The advantages of a Scotch boiler and of all internally 
fired boilers are: (1) minimum radiation losses; (2) requires no set- 
ting; (3) no leakage of cool air into the furnace as sometimes occurs 
through cracks or porous brickwork of other types; (4) large steam- 
ing capacity for the space occupied. The circulation, however, is not 
always positive and the water below the furnace maj r be considerably 



94 



STEAM POWER PLANT ENGINEERING 




02 



BOILERS 



95 



below the average or normal temperature, giving rise to unequal expan- 
sion and contraction which may cause leakage. The boiler proper is 
relatively costly, but this is offset to some extent by the absence of 
setting. 

64. Robb-Mumford Boiler. — Fig. 35 shows a section through a 
Robb-Mumford boiler, which is a modification of the Scotch-marine 
and of -the horizontal tubular type. It consists of two cylindrical 
shells, the lower one containing a round furnace and tubes and the 
upper one forming the steam drum, the two being connected by two 
necks. The lower shell has an incline of about one inch per foot from 
the horizontal, for the purpose of promoting circulation and draft, 




/ W///////M/////M///my/^^ 

Fig. 35. Robb-Mumford Boiler. 



and also for convenience in washing out the lower shell. Combustion 
takes place in the furnace, which is surrounded entirely by water, and 
the gases pass through the tubes and return between the lower and 
upper shells (this space being inclosed by a steel casing) to the outlet 
at the front of the boiler. Mingled water and steam circulate rapidly 
up the rear neck into the steam drum, where the steam is released, the 
water passing along the upper drum towards the front of the boiler and 
down the front neck, a semi-circular baffle plate around the furnace 
causing the down-flowing water to circulate to the lowest part of the 
lower shell under the furnace. The outer casing, which incloses the 
space between the lower and upper shells, including the rear smoke 
box and the smoke outlet, is constructed of steel plate, with angle-iron 
stiffeners, the various sections being bolted together for convenient 
removal. The inside of the steel case, including the rear smoke cham- 



96 STEAM POWER PLANT ENGINEERING 

ber, is lined with asbestos air-cell blocks fitted in between the angle-iron 
stiff eners. The top of the upper drum and bottom of the lower shell 
are also covered with non-conducting material after the boiler is erected. 
Owing to the fact that steam and water spaces are divided between two 
cylindrical shells, the thickness of plates is not so great as in the Scotch- 
marine or horizontal return tubular types; and the rear chamber of the 
marine boiler is avoided. 

The chief claim for this type of boiler is compactness. A battery 
of five 200-horse-power units occupies a floor space of but 33 
feet in width by 20 feet in depth and 12.5 feet high. Each unit 
is entirely independent and may be isolated for cleaning, inspection, 
and repairs. 

65. Horizontal Return Tubular Boilers. — These are the most com- 
mon in use and are constructed in sizes up to 200 horse power. They 
are simple and inexpensive and, when properly operated, durable and 
economical. Figs. 36 to 39 show various forms of standard settings, 
and Figs. 85, 86, and 87 different " smokeless" settings. The grate is 
independent of the boiler, and the products of combustion pass beneath 
the shell to the back end, returning through the tubes to the front, 
and into the smoke connection. 

The tubes are from 3 to 4 inches in diameter and from 14 to 18 feet long, 
and are expanded into the tube sheets. The portion of the tube sheets 
not supported by the tubes is secured against bulging by suitable stays. 
Access to the interior of the boiler is obtained through manholes. The 
most convenient arrangement for inspection and cleaning is to have 
one manhole located at the top of the shell and one at the bottom of 
the front tube sheet. Return tubular boilers are made either with an 
extended front (Fig. 36) or flush front (Fig. 37). The latter costs a 
little more for brick and setting, but it is more convenient to operate 
and the boiler is less expensive. The shell may be supported by lugs 
on the brickwork as in Fig. 36 or by steel beams and hangers as in 
Fig. 38. The latter construction permits the brickwork and shell to 
expand or contract independently, and settling of the brickwork does 
not affect the boiler alignment. With the side bracket support, the 
front lugs usually rest directly on iron or steel plates embedded in the 
brickwork, and the back lugs on rollers, to permit free expansion* and 
contraction. The brackets are long enough to rest upon the outside 
wall, so that the inside brick lining can be renewed without disturbing 
the setting. The distance between the rear tube sheet and wall should 
be about 16 inches for boilers less than 60 inches in diameter and from 
20 to 24 inches for larger ones. The distance between grate and boiler 
shell should not be less than 28 inches for anthracite coal and 36 inches 



BOILERS 



97 




£ 




98 



STEAM POWER PLANT ENGINEERING 







BOILERS 



99 



for bituminous coal.* The greater this distance the more complete 
the combustion, since the gases will have a better opportunity for com- 
bining with the air before coming into contact with the comparatively 
cool surfaces of the shell. The shell should be slightly inclined toward 
the blow-off end so as to drain freely. 




Fig. 38. Return Tubular Boiler Setting — Steel Beam Suspension. 

The vertical distance between the bridge wall and shell is usually 
between 10 and 12 inches. The lower part of the combustion chamber 
behind the bridge wall may be filled with earth and paved with common 
brick as in Fig. 39 or left empty as in Fig. 37. The shape of the bridge 
walls whether curved to conform to the shell or flat appears to have 
little influence on the economy. 

The side and end walls are ordinarily constructed of common brick 
with an inner lining of fire brick, and may be solid as in Fig. 37 or 
double with air spaces as in Fig. 36. f The latter construction is pref- 
erable and permits the inner and outer walls to expand independently 
without cracking and settling. The side walls are braced by five pairs 
of buckstaves, with through rods under the paving and over the tops 
of the boilers. 

* For smokeless combustion the setting must be modified. See furnaces illus- 
trated and described in paragraph 97. 

f See "The Flow of Heat through Furnace Walls," Bulletin No. 8, U. S. Bureau 
of Mines, 1911. 



100 



STEAM POWER PLANT ENGINEERING 




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BOILERS 



101 



The connection between the rear wall and the shell is a source of 
more or less trouble on account of the expansion and contraction of the 
boiler. Cast-iron supports of T section supporting a fire-brick arch 
are usually employed as illustrated in Fig. 40, the clearance between 
the arch and the shell being sufficient to allow the necessary expansion. 

Fig. 41 shows the common method of resting one end of the arch 
supports on the rear wall and the other end on an angle iron riveted to 
the boiler. 

The products of combustion are sometimes carried over the top of 
the boiler as shown in Fig. 39. This tends to superheat the steam, but 
the advantage gained is probably offset considerably by the extra cost 
of the setting and the accumulation of soot on the top of the shell. The 
arrangement is not common. 




Fig. 40. Furnace Arch Bars. 



Fig. 41. 



Back Connection made with 
Cast-iron Plate. 



The steam connection is naturally made to the highest point in the 
boiler shell. Frequently a steam dome, to which the steam nozzle is 
connected, is provided as in Fig. 37. The function of the steam dome 
is to increase the steam space so as to permit the collection of dry 
steam at a point high above the water level. If a boiler is too small for 
its work and is forced far above its rating a steam dome is probably an 
advantage, though its use is less common now than formerly, since a 
properly designed boiler insures ample steam space without one. A 
dry pipe inside the boiler above the water line as in Fig. 34 or 35 is 
commonly used to guard against priming where the nozzle is connected 
to the shell. 

For low pressures and small powers the return tubular boiler has the 
advantage of affording a large heating surface in a small space and large 
overload capacity. It requires little overhead room and its first cost 
is low. On the other hand the interior is difficult of access for purposes 



102 



STEAM POWER PLANT ENGINEERING 



of cleaning and inspection. Boilers of this type are constructed in 
various sizes ranging from a 36" X 8', rated at 15 horse power, to a 
96" X 21', rated at 400 horse power, though sizes above 200 horse 
power are exceptional. The working pressure seldom exceeds 150 
pounds per square inch. 

66. Lyons Boiler. — The standard externally fired return tubular 
boiler is limited in size since the danger from overheating the shell 
directly over the fire bed increases rapidly with the increase in thick- 
ness of the plate. The Lyons boiler, a section through which is shown 




Fig. 42. Lyons Boiler and Setting. 



in Fig. 42, overcomes this restriction through the addition of a bank 
of water tubes which form a roof to the furnace. These tubes protect 
the shell from the direct action of the gases and insure a positive and 
rapid circulation. They are covered with tile or split brick and form 
the equivalent of a " Dutch oven." Lyons boilers are made in various 
sizes up to 450 horse power. 

67. Sederholm Boiler. — Fig. 43 shows a longitudinal section and a 
vertical sectional elevation of a Sederholm boiler with stationary grate 
and setting. This boiler is a modification of the old "elephant" type 
so much in evidence in France. As will be seen from the illustration 
it consists essentially of a return tubular boiler fitted with four hori- 



BOILERS 



103 




104 



STEAM POWER PLANT ENGINEERING 



zontal water drums extending the full width of the setting. The 
drums form a roof for the furnace and protect the bottom of the shell 
from direct flame impingement. A large portion of the impurities in 
the feed water is precipitated in the drums from which it is readily 







discharged. A number of Sederholm boilers equipped with chain- 
grate stokers are installed in the power plant of the Commercial National 
Bank Building, Chicago, and are giving excellent results. 

68. Babcock and Wilcox Boiler. — Fig. 44 shows a longitudinal section 
through a Babcock and Wilcox boiler, illustrating a typical horizontal 



BOILERS 



105 



water-tube type. The tubes, usually 4 inches in diameter and 18 feet 
in length, are arranged in vertical and horizontal rows and are expanded 
into pressed-steel headers. Two vertical rows are fitted to each header 
and are " staggered" as shown in Fig. 45. The headers are connected 
with the steam drum by short tubes expanded into bored holes. Each 
tube is accessible for cleaning through openings closed by covers with 
ground joints held in place by wrought-iron clamps and bolts. The 
tubes are inclined at an angle of about 22 degrees with the horizontal. 



Fig. 45.- Details of 
Header — Babcock 
and Wilcox Boiler. 




Fig. 46. Front Section — Babcock and Wilcox Boiler. 



The rear headers are connected at the bottom to a cast-iron mud drum. 
The steam drum is horizontal and the headers are arranged either ver- 
tically or at right angles to the tubes. The boiler is supported by steel 
girders resting on suitable columns independent of the brick setting. 
The grate is placed under the higher ends of the tubes, the products of 
combustion passing at right angles to the tubes and being deflected 
back and forth by fire-tile baffles. The feed water enters the front of 
the steam drum as shown in Fig. 46. A rapid circulation is effected by 
the difference in density between the solid column of water in the rear 
header and the mixed steam and water in the front one. B. & W. 



106 



STEAM POWER PLANT ENGINEERING 



boilers under 150 horse power have but one steam drum, and the larger 
sizes have two. The number of tubes varies with the size of boiler, 




ranging from 6 wide and 9 high in the 100-horse-power boiler to 
14 high and 18 wide in the 500-horse-power boilers. 

69. Heine Boiler. — Fig. 47 shows a longitudinal section through a 
Heine horizontal water-tube boiler. This boiler differs from the B. & W. 



BOILERS 107 

boiler in that the tubes are expanded into a single large header con- 
structed of boiler steel. The drum and tubes are parallel with each 
other and inclined about 22 degrees with the horizontal. The feed 
water enters at the front of the steam drum and flows into the mud 
drum, from which it passes to the rear header. Steam is taken from 
the front of the steam drum and is partially freed from moisture by the 
dry pipe A. A baffle over the front header prevents an excess of water 
from being carried into the dry pipe. As the rear header forms one large 
chamber, no additional mud drum is necessary and the sediment is 
blown off from the bottom by the blow-off cock. The circulation is 
somewhat freer than in the B. & W. boiler on account of the large 
sectional area through the headers. 

70. Wickes Boiler. — Fig. 48 shows a section through a Wickes 
vertical boiler, illustrating the vertical water-tube type. The steam 
drum and water drum are arranged one directly above the other. The 
tubes are expanded and rolled into both tube sheets and are divided 
into two sections by fire-brick tile. The water line in the steam drum 
is carried about two feet above the tube sheet, leaving a space of five 
feet between water line and top of the drum. This affords a large 
steam space and disengagement surface. Feed water is introduced 
into the steam drum below the water line and flows downward through 
the tubes of the second compartment. The boiler is supported by four 
brackets riveted to the shell of the bottom drum and is independent 
of the setting. The entire boiler is inclosed in brickwork and is com- 
pletely surrounded by the products of combustion. The upper part 
of the steam drum acts as a superheating surface and tends to dry 
the steam. Wickes boilers are simple in design, easy to inspect and 
clean, low in first cost, and comparable in efficiency with any water- 
tube type of boiler. 

71. Parker Boiler. — Fig. 49 shows a longitudinal sectional elevation 
and an end sectional elevation of a 1200-horse-power Parker down- 
flow boiler with double-ended setting. This type of boiler is finding 
much favor with engineers for central stations where large units are 
desired. The Parker boiler differs from the conventional horizontal 
water-tube boiler principally in circulation and flexibility. 

Feed water is pumped into the economizer or feed element (1), Fig. 
49, at 0, 0, and flows downward through a series of tubes, discharging 
finally into the drum through an upcast H. In a large unit, as illus- 
trated here, there are two feed elements and two drums. The circula- 
tion in the feed element is indicated by solid lines and arrow points at 
the left of the end sectional elevation, the tubes having been omitted 
from the drawing for the sake of clearness. 



108 



STEAM POWER PLANT ENGINEERING 



The intermediate elements (2) take their water supply from the bottom 
of the drum through a cross-box V, the circulation being downward, as 
indicated by arrow points, through four tube wide elements, and finally 
discharge it through an upcast X into the steam space of the drum. 




ill 



1 1 i 1 



ill 



I 1 



on 




Fig. 48. Wickes Vertical Water-tube Boiler. 

Each element has a " down-comer" and an upcast. In the smaller- 
sized boilers the intermediate elements are omitted. 

The evaporator elements (3) take their water supply from the bottom 
of the drum at Y, the circulation being downwards through two tube 
wide elements, and finally discharge it into the drum at U. The last 
two passes of the water are through the two bottom tubes of each 



J 



BOILERS 



109 




110 



STEAM POWER PLANT ENGINEERING 



element, thus assuring dry steam without the use of dry pipes. To 
prevent reversal of flow each element is fitted with a check valve at the 
admission end. Each drum is equipped with a diaphragm, as indicated, 
separating the steam and water spaces, thus insuring against foaming 
and priming. 

Saturated steam is taken from the drum at A and passes by way of 
B to C, where it enters the superheater S. The superheated steam 
leaves the superheater at D and passes by way of E and R to the storage 
drum N, finally leaving the boiler at G. The superheater is designed 
to maintain an approximately constant degree of superheat for all 

variations in load. 

All tubes are connect- 
ed by malleable-iron 
junction boxes the in- 
terior of each tube being 
accessible through hand 
holes placed opposite 
the end of each tube. 
The hand-hole cover 
plates are on the inside 
of the box and have 
conical ground joints, 
thus dispensing with 
gaskets. 

FlG - 50 - The Parker boiler is 

built single or double ended, with or without superheater, and in sizes 
ranging from 50-horse-power to 2500-horse-power standard rating. 

72. Stirling Boiler. — Fig. 51 shows a longitudinal section through a 
Stirling water-tube boiler, which differs considerably from the types 
just described. Three horizontal steam drums and one horizontal mud 
drum are connected by a series of inclined tubes. The tubes are bent 
at the ends to permit them to enter the drums radially. Short tubes 
connect the steam spaces of all the upper drums and also the water 
spaces of the front and middle drums. Suitably disposed fire-tile 
baffles between the banks of tubes direct the gases in their proper 
course. The boiler is supported on a structural steel framework in- 
dependent of the setting. The feed water enters the rear upper drum, 
which is the cooler part of the boiler, and flows to the bottom or mud 
drum, where it is heated to such an extent that many of the impurities 
are precipitated. There is a rapid circulation up the front bank of 
tubes to the front drum, across to the middle drum, and thence down 





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SIZE? OF 


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Heine-1 

Parker-20,21 

Hormsby-18,19 








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600 



1000 1500 

Capacity of Battery Boiler H.P. 



2000 



BOILERS 111 

the middle bank of tubes to the mud drum. The interior of the drums 
is accessible for cleaning by manholes located in the ends. The Stirling 
furnace is distinctive in design. A fire-brick arch is sprung over the 
grates immediately in front of the first bank of tubes. The large tri- 
angular space between boiler front, tubes, and mud drum forms the 
combustion chamber. Stirling boilers are somewhat lower in first 
cost than other types of water-tube boilers on account of the absence 



r Fig. 51. Stirling Boiler and Setting. 

of numerous hand holes and the like which are necessary in the hori- 
zontal type. 

Fig. 52 gives a sectional view through the boiler and setting of a 
2365-horse-power Stirling boiler equipped with Taylor stokers as in- 
stalled at the Delray station of the Detroit Edison Company. Five 
boilers are now in operation and it is planned to eventually install 
ten. Though rated at 2365 boiler horse power they are capable of 
carrying continuously a load equivalent to 6000 kilowatts with a max- 
imum of 8000 kilowatts. The overall dimensions of the boiler and 



112 



STEAM POWER PLANT ENGINEERING 







Fig. 52. 2365-horse-power Stirling Boiler — Delray Station, Detroit Edison Company. 



BOILERS 



113 



setting are shown in the illustration. Each unit contains 23,654 square 
feet of effective heating surface and is provided with superheaters for 
supplying steam at 150 degrees superheat. Table 31 gives a resume 
of the principal results obtained from tests of these units with Roney 
and Taylor stokers. The grate surface per boiler for the Roney stoker 




Fig. 53. Bigelow-Hornsby Boiler and Setting. 

is 446 square feet and for the Taylor stoker 405 square feet, thus giving 
as ratios of grate surface to heating surface 1 : 53 and 1 : 58.5 respectively. 
For a complete description of these tests see Jour. A.S.M.E., Nov., 
1911, p. 1439. 

73. The Bigelow-Hornsby Boiler. — Fig. 53 shows a vertical section 
through a Bigelow-Hornsby boiler equipped with Foster superheater and 
Taylor stoker. This boiler is of the vertical water-tube type and is 
made up of a number of cylindrical elements, each element comprising 



114 STEAM POWER PLANT ENGINEERING 

an upper and lower drum connected by straight tubes. The two front 
elements are inclined over the furnace at an angle of about 68 degrees, 
and the two rear elements are vertical. The upper drums of the ele- 
ments are connected to a horizontal main steam drum by flexible tubing 
as indicated. Four elements constitute a section with an effective 
heating surface of 1250 square feet. Any number of sections may be 
connected together forming units of from 250 to 2500 boiler horse power 
or more. All parts, both external and internal, are readily accessible. 
Feed water enters the top drum of the rear elements and passes twice 
the length of the tubes before entering into the general circulation. 
This arrangement permits a considerable portion of the impurities in 
the water to be precipitated in the rear drum from which they are 
readily discharged. By the time the water reaches the front of the 
boiler directly over the furnace, where the heat transmission is the most 
intense, the scale-forming elements have been practically eliminated. 
The particular features of this boiler lie in the great extent of heating 
surface exposed to radiant heat and the height and volume of the com- 
bustion chamber. Bigelow boilers are productive of high economy 
and are readily forced to twice their rated capacity with little decrease 
in over-all efficiency. The most notable installation of Bigelow boilers 
in this country is at the power plant of the Hartford Electric Light & 
Power Company, Hartford, Conn., where two 1250- and one 2500-boiler- 
horse-power units are installed. The latter is the largest single boiler 
and setting in the world at this writing (Feb., 1912). 

74. Unit of Evaporation. — The performance of a boiler and furnace 
may be expressed in terms of the weight of water evaporated per hour 
per square foot of heating surface or of the weight evaporated per 
pound of fuel. To reduce all performances to an equal basis so as to 
facilitate comparison the evaporation under actual conditions is con- 
veniently referred to the equivalent evaporation from a feed-water 
temperature of 212 degrees F. to steam at atmospheric pressure. The 
heat required to evaporate one pound of feed water at a temperature 
of 212 degrees F. into steam of the same temperature, or "from and 
at 212 degrees'' as it is commonly called, is 970.4 B.t.u. The ratio of 
the heat necessary to evaporate one pound of water under actual con- 
ditions of feed temperature and steam pressure to the heat required 
to evaporate one pound from and at 212 degrees is called the factor of 
evaporation. Thus, for dry steam, 

* For most practical purposes q 2 may be taken as t — 32, in which t = tem- 
perature of the feed water, degrees F. 



BOILERS 



115 



in which 

F = factor of evaporation, 

X = total heat of one pound of steam at observed pressure above 

32 degrees F., 
q 2 = total heat of one pound of feed water above 32 degrees F. 

If the steam is wet, 

X = xr + q, (28) 

in which 

x = the quality of the steam, 

r = latent heat of evaporation at observed pressure, 

q = heat in liquid at observed pressure. 

If the steam is superheated, 

X = r + q + a, (29) 

in which 

C = the mean specific heat of the superheated steam, 
t 8 = the degree of superheat, degrees F. 




Furnace Walla 



Heat from Hot ' 
'.Moving Gases i 



2500" 



gSSSEKl Convection ) Boiler 
I Water 

1 

4 



75. Heat Transmission. — Fig. 54 shows a section through a boiler- 
heating plate and serves to illustrate the accepted theory of heat trans- 
mission. The outer surface of the ws, 



plate is covered with a thin layer 
of soot and a film of gas, and the 
inner surface is similarly protected 
by a layer of scale and a film of 
steam and water. It is, therefore, 
reasonable to assume that the dry 
surface of the plate is located 
somewhere within the film of gas, 
and the wet surface within the film 
of water and steam. 

The heat is imparted to the dry 
surface by: (1) radiation from the 
hot fuel bed and furnace walls, 
and by (2) convection from the 
moving furnace gases. The heat 
is transferred through the boiler 
plate and its coatings purely by 
conduction. The final transfer 
from the wet surface to the boiler 
is mainly by convection. 



2000- 



1000- 



<500- 




A~Average Temperature of.Moving Gases'. 
B= Average Temperature of Dry Surface. 
C = Average Temperature of Wet Surface. 
D =Temperature of Water m Boiler. 

Fig. 54. — Heat Transmission through 
Boiler Plate. 



Radiation depends on the temperature, and according to the law of 
Stefan and Boltzmann is approximately proportional to the difference 



116 STEAM POWER PLANT ENGINEERING 

between the fourth power of the absolute temperature of the fuel bed 
and furnace walls and the temperature of the dry surface of the heat- 
ing plate. According to this law the heat transmitted by radiation 
increases rapidly with the increase in furnace temperature. In the 
ordinary boiler and setting the surface exposed to radiation is only a 
small portion of the total heating surface, and, since in well-operated 
furnaces the temperature of the furnace cannot be increased materially 
on account of practical considerations, there is little hope of increasing 
the capacity of a boiler by increasing the furnace temperature. The 
extent of heating surface exposed to radiation, however, may be greatly 
increased. Many authorities are of the opinion that the boiler of the 
future will depend largely upon radiation. 

The heat imparted to a boiler plate by convection may be determined 
by the following equation (Prof. Perry, " The Steam Engine," 1906 
Ed., p. 588): 

H = C{h- t 2 ) vol, (30) 

in which 

H = B.t.u. transferred per hour per sq. ft. of heating surface, 

C = a coefficient determined by experiment, 

t\ = temperature of the moving gases, degrees F., 

t 2 = temperature of the dry plate surface, degrees F., 

v = velocity of the gases, feet per second, 

d = density of the gases, pounds per cubic foot. 

Professor Nicholson gives the following modifications of formula 
(30) as applied to boiler tubes or flues (Engr. Lond., Feb. 19, 1908): 

*-|.355 + ffii( 1 + s)! ft -*>'* (31) 

in which 

t = mean film temperature, 
m = hydraulic radius = area of tube in square inches -5- perimeter 
of the tube in inches; other notations as in (30). 

Both equations are based upon the same general law except that the 
latter gives a means of determining coefficient C in terms of the mean 
film temperature and the dimensions of the flues or tubes. 

An examination of equation (30) shows that for a given set of condi- 
tions the heat imparted by convection to a unit of dry surface of heat- 
ing plate varies directly as the difference between the temperature of 
the hot gases and that of the dry surface and directly as the velocity 
and density of the gases. However, the density of the gases drops 
with the rise of temperature, and increase in furnace temperature does 
not necessarily imply increase in heat impartation. It is the utiliza- 



BOILERS 



117 



tioiTof the velocity factor, then, which offers a possibility of increasing 
boiler capacity. 

Experiments by Professor Nicholson and the U. S. Geological Survey 
show that by establishing a powerful scrubbing action between the 
gases and the boiler plate the protecting film of gas is torn off as rapidly 



13 



12 



11 



10 



U& 9 

.firs 

tn o 

£« 8 



56 

W 

£2 



as it is formed and new portions 
of the hot gases are brought into 
contact with the plate, thereby 
greatly increasing the rate of 
heat transmission. Similarly, the 
faster the circulation of the water 
the greater will be the scrubbing 
action tending to remove the 
bubbles of steam from the wet 
surface and the more rapid will be 
the transfer from the plate to the 
boiler water. 

The resistance of the metal 
itself is so small that it may be 
neglected in calculating the heat 
transmission, and it may be 
logically assumed that the plate 
will take care of all the heat that 
reaches its dry surface. 

Professor Nicholson found that 
by filling up the flue of a Cornish 
boiler with an internal water 
vessel, leaving an annular space 
of only 1 inch around the latter, 
an evaporation eight times the 
ordinary rate was effected at a 
flow of gases 330 feet per second 
(8 to 10 times the average flow). 
The fan for creating the draft 
consumed about 4J per cent of 
the total power. 

The conclusion is that the 
heating surface for a given evaporation 
be reduced as much as 90 per cent 



£ C <B O 

111 2 



a i 



m 



-*m 



3fc 



ill 



m 



£t«3 , 
,OgWj3 

> 0+J &© 4 

100 200 300 

Per cent, of Rated Capacity Developed by Boiler. 

Fig. 55. Influence of Draft on the Capacity 
of a Normand Water-tube Boiler on the 
U. S. Torpedo Boat "Biddle." 



i% 



1 



/- 



m 



at the present rating may 

for the same output, with 

a corresponding reduction in the size, cost, and space requirements, 

or with a given heating surface of standard rating the output 

may be enormously increased; also the increase in power necessary 



118 STEAM POWER PLANT ENGINEERING 

to create the draft is by no means comparable with the advantages 
gained. 

The modern locomotive boiler is the nearest approach to these con- 
ditions in practice. Here a powerful draft forces the heated gases 
through small tubes at a very high velocity and an enormous evapo- 
ration is effected with a comparatively small heating surface. See 
Fig. 55 for influence of draft on the capacity of a torpedo-boat boiler 
(Power and Engr., May 24, 1910). 

These principles have been applied to a limited extent to stationary 
boilers already installed by making the gas passages smaller as com- 
pared to the length by means of suitable baffles (Fig. 49) and by forc- 
ing larger weights of gas through the boiler, either by forced draft or 
by increasing the grate area (Fig. 52). 

In a general sense when the capacity of a boiler is doubled or tripled 
the over-all efficiency of the whole steam-generating apparatus drops, 
but the advantage gained usually offsets the loss in fuel economy. A 
close examination of the results, however, will show that the loss in 
efficiency is due more to low furnace efficiency than to inability of the 
boiler to absorb the heat generated. 

In view of recent experiments it is not unlikely that within the next 
ten years boilers will be constructed capable of developing a boiler horse 
power with two or three square feet of heating surface instead of ten 
square feet, as at present, and with high over-all efficiency. (See 
Figs. 57 and 58.) 

Heat Transmission in Boilers, Kreisinger and Ray: Power and Engr., June 29, 
1909, p. 1144; Bulletin No. 18 U. S. Bureau of Mines, 1912; Journ. West Soc. 
Engrs., Sept. 18, 1907; Am. Inst. Elect. Engrs., Dec. 13, 1907. 

Heat Transfer and Future Boiler Practice: A. H. Allen, Power and Engr., Sept. 
21, 1909, p. 482; Engng., Lond., Feb. 19, 1908. 

The Heat of Fuels and Furnace Efficiency: W. D. Ennis, Power and Engr., July 
14, 1908, p. 50. 

A Study in Heat Transmission (The Transmission of Heat to Water in Tubes as 
Affected by the Velocity of the Water), J. K. Clement and C. M. Garland, Univ. 
of 111. Bulletin No. 40, Sept. 27, 1909; Power & Engr., Feb. 7, 1911, p. 222. 

76. Heating Surface. — All parts of the boiler shell, flues, or tubes 
which are covered by water and exposed to hot gases constitute the 
heating surface. Any surface having steam on one side and exposed 
to hot gases on the other is superheating surface. According to the 
recommendations of the American Society of Mechanical Engineers, 
the side next to the gases is to be used in measuring the extent of the 
heating surface. Thus measurements are made of the inside area of 
fire tubes and the outside area of water tubes. The heating surface in 



BOILERS 119 

a boiler under average conditions of good practice is most efficient when 
the heated gases leave the uptake at a temperature of 100 to 200 degrees 
F. above that of the steam. Each square foot of heating surface is 
capable of transmitting a certain amount of heat, depending upon the 
conductivity of the material, the character of the surface, the temperature 
difference between the gas and the water, the location and arrangement of 
the tubes, the density of the gas, the velocity of the gas, and the time allowed 
for transmission of the heat. 

Thus one square foot of heating surface in the first pass of a water- 
tube boiler immediately over the incandescent mass of fuel may evapo- 
rate as high as 50 pounds of water per hour from and at 212 degrees F., 
whereas the same extent of surface close to the breeching evaporates 
less than one pound per hour. Because of this extreme variation it is 
convenient to assume a uniform heat transmission for the entire surface 
which will give the same total evaporation as that actually obtained. 
For maximum economy under average conditions of operation this gives a 
mean evaporation of 3 to 3.5 pounds of water per square foot per hour 
from and at 212 degrees F., which is equivalent to allowing 10 to 12 
square feet per boiler horse power. By providing a large combustion 
chamber, increasing the extent of the first pass or the equivalent and 
by carrying a very thick bed of fuel a mean evaporation of 7 pounds 
per square foot per hour has been maintained with high economy. 
This corresponds to 5 square feet of heating surface per boiler horse 
power. 

The maximum evaporation is limited only by the amount of coal which 
can be burned upon the grate. For example, a mean evaporation as high 
as 20 pounds* per square foot per hour has been effected in torpedo- 
boat practice, under intense forced draft, and 12 pounds per square 
foot per hour is not unusual in locomotive work. Such extreme, high 
rates of evaporation, however, are invariably obtained at the expense 
of fuel economy. In the very latest central stations the boiler and 
settings are proportioned to operate at 100 per cent above standard 
rating with high over-all efficiency and at 200 per cent above rating 
with only a small drop in efficiency, but such results are not obtainable 
in the ordinary everyday boiler and setting. 

Builders of return tubular and vertical fire-tube boilers allow from 11 
to 12 square feet of heating surface per horse power; water-tube boilers 
are rated at 10 square feet per horse power, and Scotch-marine 
boilers at 8 square feet per horse power. 

See, also, paragraph 81, Effect of Capacity on Efficiency. 

* Eng. Mag., Jan., 1912, p. 504. 



120 



STEAM POWER PLANT ENGINEERING 



The following table shows approximately the relation between boiler 
horse power and heating surface for different rates of evaporation: 

EVAPORATION FROM AND AT 212 DEGREES F. PER SQUARE FOOT PER HOUR. 



2 


2.5 


3.0 


3.5 


4 


5 


6 7 


8 


9 


10 


SQUARE FEET HEATING SURFACE REQUIRED PER HORSE POWER. 


17.3 


13.8 


11.5 


9.8 


8.6 


6.8 


5.8 


4.9 


4.3 


3.8 


3.5 



Efficiency of Boiler Heating Surface: Trans. A.S.M.E., 18-328, 19-571. Kent, 
"Steam Boiler Economy" (John Wiley & Sons), Chapter IX. 

The Nature of True Boiler Efficiency: Jour. West. Soc. Engrs., Sept. 18, 1907. 
Heat Transference through Heating Surface: Engineering, 77-1. 

77. The Horse Power of a Boiler. — A boiler horse power is equivalent 
to the evaporation of 34.5 pounds of water per hour from a temper- 
ature of 212 degrees F. to steam at atmospheric pressure. This corre- 
sponds to 33,479 B.t.u. per hour. Since the power from steam is 
developed in the engine and the boiler itself does no work, the above 
measure of capacity is merely conventional. Thus one boiler horse 
power will furnish sufficient steam to develop about three actual horse 
power in the best compound condensing engine, but only one-half 
horse power in a small non-condensing engine. Boilers should be 
purchased on the basis of heating surface and not on the horse-power 
rating, since one bidder may offer a boiler with, say, 5 square feet of 
heating surface per horse power and another with 10 square feet, both 
being capable of the required evaporation, but the one with the small 
heating surface (which will, of course, be the cheaper boiler) will have 
considerably less reserve capacity. Manufacturers ordinarily rate their 
boilers on the basis of from 10 to 12 square feet of heating surface per horse 
power, and the power assigned is called the builder's rating. As this 
practice is not uniform, bids and contracts should always specif} r the 
amount of heating surface to be furnished. According to the recom- 
mendations of the American Society of Mechanical Engineers, "A boiler 
rated at any stated capacity should develop that capacity when using 
the best coal ordinarily sold in the market where the boiler is located, 
when fired by an ordinary fireman, without forcing the fires, while 
exhibiting good economy. And, further, the boiler should develop at 
least one-third more than stated capacity when using the same fuel 
and operated by the same fireman, the full draft being employed and 
the fires being crowded; the available draft at the damper, unless other- 
wise understood, being not less than one-half-inch water column. 



BOILERS 121 

In determining the boiler horse power required for a given engine 
horse power it is convenient to estimate the steam consumption of 
the engine under actual conditions and then ascertain the equivalent 
evaporation from and at 212 degrees F. For example, assume a single 
non-condensing engine developing 20 horse power to use 50 pounds of 
steam per horse-power hour, or 1000 pounds steam per hour; steam 
pressure, 80 pounds per square inch; feed- water temperature 120 degrees 
F. Required the boiler horse power necessary to furnish this quantity 
of steam. 

From equation (27), the factor of evaporation is 

„ \-<?2 1185.3 - 87.91 

F = ~970A ~ 970^ " L13L 

One thousand pounds of steam under the given conditions are therefore 
equivalent to 1000 X 1.131 = 1131 pounds from and at 212 degrees F. 

The boiler horse power necessary to furnish steam for the 20-horse- 
power engine will be 

1131 

Boiler horse power = ^j-- = 32.8. 

Example: A 15,000-kilowatt steam turbine and auxiliaries require 
14.7 pounds of steam per kilowatt-hour at rated load; steam pressure 
200 pounds per square-inch gauge; superheat 150 degrees F.; feed- 
water temperature, 179 degrees F. 

Required the boiler horse power necessary to furnish this quantity 
of steam. 

The heat furnished to the turbine and auxiliaries per kilowatt-hour 
is 

w\\ + C p t, - q 2 } = 14.7 { 1199.2 + 0.57 X 150 - 146.881 
= 16,724 B.t.u., 

„ ., , 15,000 X 16,724 _„_ , N 

Boiler horse power = — — = 7500 (approx.). 

00,4/ y 

For forced capacity of boilers, see Table 30. 

78. Grate Surface. — The amount of fuel which can be burned per 
hour limits the amount of water evaporated per unit of time and de- 
pends upon the extent and nature of the grate surface, the character 
of the fuel and the draft. In locomotive and torpedo-boat practice 
space limitations necessitate the use of small grates and the rate of 
combustion is primarily a direct function of the draft. In stationary 
practice there is a wide permissible range in proportioning the grate 
surface since a given rate of combustion may be effected with large 
grate surface and light draft or with small grate surface and strong 



122 STEAM POWER PLANT ENGINEERING 

draft. In a general sense the best results are obtained with a small 
grate and a high rate of combustion, but in the majority of installations 
the draft is comparatively feeble and a liberal grate area is necessary. 
So much depends upon the grade and size of the fuel that general rules 
for proportioning the grate surface are apt to lead to serious error. A 
liberal allowance of grate surface is desirable for hand-fired furnaces 
with natural draft, particularly if the ash is easily fusible, tending to 
choke the grate, but with forced draft and automatic stokers the best 
results are obtained with a thick fire and small grate surface. The 
relation between draft and rate of combustion for various sizes and 
kinds of coals is shown in Fig. 155. 

A number of boiler tests made by Barrus (" Boiler Tests") showed 
that the best economy with anthracite coal, hand-fired, was obtained 
with an average ratio of grate surface to heating surface of 1 to 36, and 
at a rate of combustion of approximately 12 pounds of coal per square 
foot of grate surface per hour. In these tests a variation in grate and 
heating-surface ratio of 1 to 36 up to 1 to 46 gave practically no differ- 
ence in economy. With bituminous coal the tests showed that an average 
ratio of 1 to 45 gave the best results and at a rate of combustion of 
24 pounds of coal per square foot of grate surface per hour. 

Tests made by Christie (Trans. A.S.M.E., 19-330) gave an average 
combustion of 13 pounds of anthracite per square foot of grate per hour 
for maximum efficiency and 24 pounds of bituminous. 

Current central-station practice gives average rates of combustion as 
follows : 

POUNDS OF COAL BURNED PER HOUR PER SQUARE FOOT OF GRATE SURFACE. 

(Natural Draft.) 

Anthracite, nut 15-20 Semi-bituminous, run of mine 20-30 

Anthracite, pea 12-18 Semi-bituminous, screenings . 20-30 

Anthracite, buckwheat No. 1 . . 8-12 Bituminous, run of mine .... 20-45 

Anthracite, buckwheat No. 3. . 6-10 Bituminous, screened nut. . . . 20-40 

Semi-anthracite, run of mine . . 18-25 Bituminous, screenings 20-35 

Semi-anthracite, screenings. . . . 12-22 Bituminous, slack 18-30 

With forced draft these rates of combustion may be greatly increased. 
Some idea of the extreme rate of combustion in modern locomotive 
practice may be obtained from the following figures which give the 
pounds of coal burned per hour per square foot of grate surface for 
various conditions of operation: 

Maximum rate 200 Average rate 80 

Very high rate 150 Economical rate 60 

Average high rate 100 Low rate 50 



BOILERS 123 

Table 26 gives the relation between heating and grate surface in a 
number of recent boiler installations using different kinds of coal, and 
is illustrative of current practice. 

In proportioning the grate surface for a proposed installation the 
principal factor considered is the character of the fuel, a study being 
made of the various fuels available, and the' one selected which gives 
the highest evaporation per dollar (all items entering into the handling 
and combustion of the fuel being considered). This information may 
usually be obtained from records of plants using the same grade of fuel 
and grates similar to those intended for the proposed plant. 

79. Boiler and Furnace Efficiency. — A perfect or ideal boiler and 
furnace is one which transmits to the water in the boiler the total heat 
of the fuel. In order to effect this result combustion must be complete, 
there must be no radiation or leakage losses and the products of com- 
bustion must be discharged at the initial temperature of the fuel. No 
commercial form of steam boiler can fulfill these conditions, hence the 
amount of heat absorbed by the boiler will always be less than the 
calorific value of the fuel. 

The efficiency of the boiler and grate, and that of the boiler alone as 
recommended by the A.S.M.E., Rules for Conducting Boiler Tests, 
(Jour. A.S.M.E., Nov., 1912) may be expressed as 

Heat absorbed by the boiler 

,-,- • c -U -i a * P er Pound of coal as fired /0 _ N 

Efficiency of boiler and grate = ~~, — ^ ; ? r , (32) 

Calorific value of one pound 

of coal as fired 
and that of the boiler alone, 

Heat absorbed by the boiler per pound 

-m~. . - , ., of combustible burned on the grate, , nn . 

Efficiency of boiler = ^-j — ^ -, s =-2- L_ (32a) 

Calorific value of one pound of com- 

bustible as fired. 

The calculation of these efficiencies is illustrated by the following 
example : 

DATA AS OBSERVED. 

Steam pressure, pounds per square inch (gauge) 151.0 

Barometer, inches of mercury 28.5 

Temperature of feed water, degrees F 161.8 

Temperature of the furnace, degrees F -. 2100.0 

Temperature of flue gases, degrees F. . 480.0 

Temperature of boiler room, degrees F 60.0 

Quality of steam, per cent 98.0 

Water apparently evaporated, pounds per hour 86,000 

Coal as fired, pounds per hour 10,000 

Refuse removed from ash pit, pounds per hour 1600 



124 STEAM POWER PLANT ENGINEERING 

COAL ANALYSIS, PER CENT OF COAL AS FIRED. 

Moisture 8 

Ash 12 

B.t.u. per pound, 11,250. 

CALCULATED DATA. 

Water apparently evaporated per pound of coal as fired, pounds = 86,000 ~ 10,000 

= 8.60. 
Factor of evaporation* = [0.98 X 856.8 + 338.2 - (161.8 - 32)1 -=- 970.4 = 1.08. 
Equivalent evaporation per pound of coal as fired, pounds = 8.6 X 1.08 = 9.288. 
Heat absorbed by the boiler per pound of coal as fired, B.t.u. = 9.288 X 970.4 = 

9,013.0. 
Efficiency of boiler and grate, per cent = (9.013 -=- 11,250) 100 = 80.11. 
Refuse in ash referred to coal as fired, per cent = (1600 -r- 10,000) 100 = 16.0. 
Combustible burned on the grate, per cent of coal as fired = 100 — (8 + 16) = 76.0. 
Equivalent evaporation per pound of combustible burned, pounds = 9.288 -f- 0.76 

= 12.221. 
Heat absorbed per pound of combustible burned, B.t.u. = 12.221 X 970.4 = 

11,860. Combustible as fired, per cent = 100 - (8 + 12) = 80.0. 
Calorific value of the combustible as fired, B.t.u. = 11,250 -f- 0.80 = 14,062. 
Efficiency of the boiler, per cent = (11,860 + 14,062) 100 = 84.34. 

The efficiency of the grate alone might be expressed as 

-r^ . e , Efficiency of boiler and grate 

Efficiency of grate = -rW • fi — n — » 

Efficiency of boiler 

which is equivalent to 

^rr . - Combustible actually burned 
Efficency of grate = Combustible fired 

For the problem cited above, 

Efficiency of grate = 100 ( ' J = 95 per cent, 

or, 

76 
Efficiency of grate = 100 -^ = 95 per cent. 

o0 

This offers a good check on the calculations. 

For oil fuel furnaces and coal furnaces equipped with stokers and 
forced draft appliances the net efficiency of the boiler and furnace may 
be taken as the boiler and furnace efficiency minus the equivalent 
heat required to feed the fuel and to create the draft. 

Since the commercial form of boiler cannot possibly absorb all of the 
heat generated by the combustion of the fuel some authorities are of 
the opinion that the "true" efficiency of the boiler should be denned 
as the ratio of the heat absorbed to that actually available. Thus the 
U. S. Geological Survey defines the heat absorbed as the difference 
between the heat generated in the furnace and that discharged into 

* See footnote, par. 74. 



BOILERS 125 

the flue, and the available heat is defined as the difference between the 
heat generated in the furnace and that discharged by the products of 
combustion at the temperature of the saturated steam. 

If w = weight of the products of combustion, pounds per hour, 
t f = temperature of the furnace, degrees F., 
t c = temperature of the flue gases, degrees F., 
i s = temperature of the saturated steam, degrees F., 
t = temperature of the boiler room, degrees F., 
Cf, c c , c s = mean specific heat of the products of combustion for tempera- 
ture ranges t to t f , t C) t 8 respectively. 
Then 

weft/ = heat generated in the furnace above t degrees F., B.t.u. per hour, 
wc c t c = heat carried away by the flue gases above t degrees F., 
wc 8 t s = heat carried away by the flue gases, if the temperature were 
lowered to t s degrees F., 

E, = the "true" boiler efficiency = WCft f ~ WCc [% 

WC/tf — WC s t a 

= c f tf - c c t m 
C/tf — c s t 3 

For most practical purposes it is sufficiently accurate to assume a 
constant value for the mean specific heats since the actual variation in- 
fluences the result but slightly. 

Assuming c/ = c c = c 3 , 

E, = % f^j ■ (34) 

If — t a 

For the problem cited above 

^ = 100 ( 21Q0-366 J = 93 ' 4perCent ' 

R. S. Hale (Trans. A.S.M.E., 20-769) gives as the efficiency of the 
furnace or combustion 

S + F 
Efficiency of furnace = — ^ — > (35) 

£1 

in which S = B.t.u. absorbed by the boiler per pound of dry coal, 
F = B.t.u. lost in the flue gases per pound of dry coal, 
H = calorific value of one pound of dry coal. 

The efficiency of the ideal or perfect steam boiler may be expressed as 

E* = ^-^- (36) 

in which H = calorific value of the coal as fired, 

/ = inherent losses as analyzed in paragraph 30, B.t.u. 
per pound of coal as fired. 



126 



STEAM POWER PLANT ENGINEERING 



The efficiency ratio or the extent to which the theoretical possi- 
bilities are realized may be expressed as 

E* = J-i (37) 

in which 

E = efficiency of the boiler and grate (A.S.M.E. code), 
E 2 = as in equation (36). 

The chief objection to the various efficiencies as defined in equations 
(34) to (36) is the difficulty of determining with any degree of accuracy 
the weight of the flue gases and the mean furnace temperature. For 
this reason commercial tests of boilers include only the efficiencies as 
recommended by the A.S.M.E. code. 

Furnace Efficiency: Joseph Harrington, Jour. W. Soc. Engrs., Sept. 23, 1912. 

TABLE 24. 

RELATION BETWEEN FUEL CONSUMPTION AND BOILER AND FURNACE 

EFFICIENCY. 

(Pounds of Fuel Burned per Boiler-Horse-Power Hour. ) 



Calorific Value 






Boiler and Furnace Efficiency. 








n f TTiipI R t n 




















\JL JL Ucl, J_> . 1 . U . 

per Pound. 


40 


45 


50 


55 


60 


65 


70 


75 


80 


85 


7,500 


11.17 


9.91 


8.94 


8.12 


7.45 


6.87 


6.37 


5.95 


5.58 


5.25 


8,000 


10.45 


9.30 


8.37 


7.60 


6.97 


6.43 


5.98 


5.58 


5.22 


4.92 


8,500 


9.84 


8.75 


7.87 


7.12 


6.56 


6.05 


5.62 


5.25 


4.97 


4.63 


9,000 


9.30 


8.25 


7.45 


6.76 


6.20 


5.72 


5.31 


4.96 


4.65 


4.36 


9,500 


8.80 


7.83 


7.05 


6.40 


5.87 


5.41 


5.02 


4.69 


4.40 


4.14 


10,000 


8.37 


7.44 


6.70 


6.09 


5.58 


5.15 


4.79 


4.46 


4.18 


3.94 


10,500 


7.98 


7.09 


6.39 


5.80 


5.86 


4.90 


4.56 


4.26 


3.99 


3.76 


11,000 


7.60 


6.79 


6.09 


5.52 


5.06 


4.67 


4.34 


4.05 


3.80 


3.59 


11,500 


7.28 


6.49 


5.83 


5.29 


4.85 


4.47 


4.16 


3.88 


3.64 


3.45 


12,000 


6.97 


6.22 


5.58 


5.06 


4.65 


4.28 


3.99 


3.72 


3.48 


3.28 


12,500 


6.69 


5.97 


5.35 


4.86 


4.46 


4.11 


3.82 


3.57 


3.34 


3.14 


13,000 


6.44 


5.74 


5.15 


4.68 


4.29 


3.96 


3.68 


3.43 


3.22 


3.02 


13,500 


6.20 


5.52 


4.96 


4.51 


4.18 


3.81 


3.54 


3.31 


3.10 


2.91 


14,000 


5.98 


5.33 


4.79 


4.35 


3.99 


3.68 


3.42 


3.19 


2.99 


2.81 


14,500 


5.77 


5.15 


4.62 


4.20 


3.84 


3.54 


3.30 


3.08 


2.88 


2.72 


15,000 


5.58 


4.96 


4.47 


4.06 


3.72 


3.43 


3.19 


2.98 


2.79 


2.64 



80. Boiler Performances. — Table 26 is compiled from a number of 
tests of different types of boilers with various types of grates and 
characters of fuel. Although some of the tests show a combined 
efficiency of boiler and grate as high as 83 per cent, such a perform- 
ance cannot be expected for continuous operation under the average 
conditions of practice. In pumping stations or in plants where there 



BOILERS 



127 



are no peak loads and the boiler may be operated under a practically 
constant set of conditions a continuous efficiency of 75 per cent has 
been realized with coal as fuel and 80 per cent with crude oil, though 
these figures are exceptional. In very large central stations, with the 
usual peak loads in the morning and evening and long banking periods, 
an average efficiency throughout the year of 65 per cent is possible, 
though a good figure is not far from 60 per cent. In large isolated 
stations with variable loads good practice gives an average of 60 per 
cent. Small stations though showing an efficiency as high as 75 per 
cent at times seldom average 50 per cent for the year. The usual 
discrepancy between efficiency as determined by special tests and 

TABLE 25. 

RELATION BETWEEN RATE OF EVAPORATION PER POUND OF FUEL AND 
BOILER AND FURNACE EFFICIENCY. 

Pounds of Water Evaporated per Hour from and at 212 Deg„ F„ per pound of FueL 



Calorific Value 








Boiler and Furnace Efficiency. 








of Fuel, 






















B.T.U. 






















per Pound. 


40 


45 


50 


55 


60 


65 


70 


75 


80 


85 


7,500 


3.09 


3.48 


3.86 


4.25 


4.64 


5.02 


5.41 


5.80 


6.18 


6.57 


8,000 


3.30 


3.71 


4.12 


4.55 


4.95 


5.36 


5.77 


6.18 


6.60 


7.01 


8,500 


3.51 


3.94 


4.38 


4.81 


5.26 


5.70 


6.14 


6.57 


7.01 


7.45 


9,000 


3.71 


4.18 


4.64 


5.10 


5.56 


6.04 


6.50 


6.96 


7.42 


7.90 


9,500 


3.92 


4.41 


4.90 


5.39 


5.88 


6.47 


6.86 


7.35 


7.85 


8.33 


10,000 


4.12 


4.64 


5.16 


5.66 


6.19 


6.70 


7.21 


7.74 


8.25 


8.76 


10,500 


4.31 


4.86 


5.40 


5.94 


6.48 


7.01 


7.55 


8.10 


8.64 


9.17 


11,000 


4.52 


5.09 


5.65 


6.22 


6.79 


7.35 


7.91 


8.48 


9.05 


9.61 


11,500 


4.74 


5.31 


5.91 


6.50 


7.10 


7.69 


8.28 


8.86 


9.45 


10.0 


12,000 


4.94 


5.55 


6.16 


6.78 


7.40 


8.01 


8.64 


9.25 


9.86 


10.5 


12,500 


5.14 


5.78 


6.42 


7.06 


7.70 


8.35 


9.00 


9.64 


10.3 


11.0 


13,000 


5.35 


6.01 


6.69 


7.35 


8.01 


8.69 


9.35 


10.0 


10.7 


11.4 


13,500 


5.56 


6.25 


6.95 


7.65 


8.34 


9.03 


9.72 


10.4 


11.1 


11.8 


14,000 


5.75 


6.48 


7.20 


7.91 


8.64 


9.35 


10.1 


10.8 


11.6 


12.2 


14,500 


5.96 


6.70 


7.45 


8.20 


8.95 


9.70 


10.5 


11.2 


12.0 


12.7 


15,000 


6.18 


6.95 


7.72 


8.50 


9.26 


10.1 


11.8 


11.6 


12.4 


13.1 



every-day operation is due to the fact that the efficiency test is usually 
conducted under ideal conditions: the boiler surfaces are cleaned, the 
rate of combustion carefully adjusted for maximum economy, and 
special attention given to the firing, whereas in actual practice these 
refinements are seldom attempted. Much depends upon the efficiency 
of the boiler-room staff, the character of furnace and fuel, draft, and 
the load factor. From the commercial standpoint the performance is 
conveniently expressed in terms of the "cost to evaporate 1000 pounds 



128 



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130 



STEAM POWER PLANT ENGINEERING 



of water from and at 212," or the " pounds of water evaporated per $1 
of coal." Table 28 gives the results of a number of tests, made at 
the Armour Glue Works, Chicago, 111., showing the cost of evaporating 
water with different grades of Illinois coal. The results were obtained 
from hand-fired Stirling boilers. 

81. Effect of Capacity on Efficiency. — In general, as the horse power 
of a boiler increases above normal capacity the over-all efficiency 
will decrease, due to the fact that the furnace and gas passages are 
ordinarily proportioned to effect an evaporation of about 3.5 pounds of 
water from and at 212 degrees F. per square foot of heating surface per 
hour at rated load, the temperature of the escaping gases being from 



0.6 






















































































§30.5 






600 B & W BOILER EQUIPPED 

WITH L TYPE GREEN CHAIN 

GRATE AT THE INTERBOROUGH 

RAPID TRANSIT CO., N.Y. 

MARCH 1912 


















































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Capacity, Boiler Horse-power 
Fig. 56. Influence of Draft on Capacity. 



150 to 200 degrees above that of the steam. To increase the rate of 
evaporation more coal must be burned per unit of time and conse- 
quently a larger volume of gas is generated. The larger the volume of 
gas the higher will be its velocity, which finally reaches a point where 
heating surface is insufficient in extent to absorb the extra heat and as 
a consequence the flue gas escapes at a higher temperature, resulting in 
lower boiler and furnace efficiency. See Fig. 57. With properly pro- 
portioned grate, furnace and gas passages a boiler may be operated at 
100 per cent above standard rating with little or no decrease in over- 
all efficiency. Figs. 56 to 62 illustrate various phases of boiler per- 
formances under different conditions of operation. These curves are 
of value simply as illustrations of the behavior in specific cases, and 
are not applicable to all types of boilers. See also Tables 30 and 32. 



BOILERS 



131 



TABLE 27. 

PRINCIPAL DATA AND RESULTS OF TESTS ON BOILER NO. 6, UNIT NO. 10, 

' FISK ST. STATION. COMMONWEALTH EDISON CO., CHICAGO. 

(B. & W. Boiler, " Standard" Setting.) 



Water-heating Surface, 5000 Sq. Ft. Superheating Surface, 914 Sq. Ft. 
Chain Grate Surface, 90 Sq. Ft. 











H.P. per 
Sq. Ft. 
Grate. 


Heat 


Total 


Super- 


Dry Coal 


Test 


Date, 


Horse 


Eff'y, 


Lost in 


Heating 


heat of 


per Sq. Ft. 


No. 


1908. 


Power. 


PerCent. 


Refuse, 


Surface 


Steam, 


G. S. 










PerCent. 


per H.P-. 


Deg. F. 


per Hour. 


2 


Mar. 9 


873 


67.4 


9.70 


2.8 


6.76 


197 


41.2 


4 


" 10 


873 


69.0 


9.52 


2.8 


6.89 


195 


39.1 


6 


" 11 


852 


67.3 


9.47 


2.8 


6.93 


189 


38.9 


8 


" 16 


836 


65.3 


9.29 


6.4 


7.06 


174 


39.5 


10 


" 17 


870 


68.8 


9.67 


5.0 


6.78 


180 


39.3 


14 


" 19 


920 


66.2 


10.22 


9.2 


6.42 


187 


43.7 


16 


" 23 


900 


69.5 


10.00 


4.0 


6.56 


181 


40.5 


18 


" 24 


916 


69.1 


10.18 


5.5 


6.44 


190 


41.6 


20 


" 26 


912 


69.2 


10.13 


4.4 


6.48 


179 


41.2 


22 


" 27 


906 


67.7 


10.07 


4.1 


6.52 


194 


42.5 


24 


" 30 


925 


69.8 


10.28 


2.8 


6.38 


179 


41.6 


26 


" 31 


894 


69.4 


9.93 


5.2 


6.60 


170 


40.6 


28 


Apr. 1 


922 


71.2 


10.24 


3.6 


6.40 


169 


40.4 


30 


" 2 


923 


71.5 


10.26 


4.6 


6.40 


173 


40.5 


32 


u 7 


914 


70.0 


10.20 


4.5 


6.46 


175 


40.9 


34 


" 8 


939 


73.8 


10.4 


3.8 


6.28 


181 


40.4 


36 


" 10 


911 


70.9 


10.1 


3.0 


6.48 


185 


40.2 


38 


" 11 


967 


70.1 


10.7 


3.0 


6.11 


192 


42.6 


40 


" 13 


995 


67.8 


11.1 


3.4 


5.93 


211 


43.6 


42 


" 14 


887 


66.8 


9.9 


•4.5 


6.65 


202 


40.8 


44 


" 27 


880 


69.5 


9.8 


5.5 


6.72 


169 


39.7 


48 


" 29 


927 


71.5 


10.3 


3.3 


6.37 


171 


40.8 


50 


" 30 


899 


70.3 


10.0 


4.2 


6.57 


171 


39.6 


52 


May 6 


886 


69.4 


9.8 


5.3 


6.67 


171 


38.2 


54 


" 7 


900 


69.1 


10.0 


4.8 


6.56 


171 


39.2 


56 


" 8 


967 


71.9 


10.7 


4.8 


6.10 


164 


40.1 


58 


" 11 


902 


70.5 


10.0 


3.3 


6.55 


163 


39.6 


60 


" 13 


875 


70.7 


9.7 


3.8 


6.74 


147 


38.3 


64 


" 14 


1102 


72.0 


12.2 


4.8 


5.35 


180 


43.2 



132 



STEAM POWER PLANT ENGINEERING 



TABLE 27. 

PRINCIPAL DATA AND RESULTS OF TESTS ON BOILER NO. 6, UNIT NO. 10, 

FISK ST. STATION. COMMONWEALTH EDISON CO., CHICAGO. 

(B. & W. Boiler, "Standard" Setting.) 

Water-heating Surface, 5000 Sq. Ft. Superheating Surface, 914 Sq. Ft. 
Chain Grate Surface, 90 Sq. Ft. 



Draft 


B.T.U. 
per Pound 
Dry Coal. 


Ash in 
Dry Coal, 
Per Cent. 


Ash in 

Refuse, 

Per Cent. 


Uptake 
Temp. 
Deg. F. 


co 2 , 

Per Cent. 


Heat Lost 
up Stack 
(Dry Gas), 
Per Cent. 


Over 
Fire. 


In 
Uptake. 


.87 


1.34 


11,634 


18.46 


82.33. 


466 


6.9 




.78 


1.25 


11,759 


16.81 


81.36 


461 


6.7 




.83 


1.25 


12,039 


16.08 


80.03 


463 


7.7 


15*6 


,94 


1.34 


11,993 


15.91 


67.42 


477 


7.6 


16.8 


.84 


1.24 


11,909 


15.71 


71.32 


475 


7.9 


16.2 


.99 


1.41 


11,768 


16.04 


63.78 


479 


8.5 


15.4 


.77 


1.17 


11,846 


16.68 


79.04 


483 


9.1 


14.0 


.81 


1.25 


11,800 


16.39 


71.98 


484 


8.3 


15.8 


.77 


1.21 


11,846 


15.51 


78.53 


486 


9.0 


14.5 


.78 


1.22 


11,659 


17.59 


80.58 


494 


9.2 


14.6 


.68 


1.28 


11,800 


16.22 


82.97 


487 


8.8 


15.1 


.70 


1.24 


11,752 


16.18 


76.84 


484 


8.8 


. 15.1 


.62 


1.21 


11,862 


15.38 


82.99 


480 


9.2 


14.1 


.58 


1.40 


11,800 


16.02 


78.37 


480 


9.1 


14.4 


.73 


1.24 


11,815 


16.84 


77.84 


494 


9.0 


14.7 


.72 


1.25 


11,659 


18.06 


82.27 


504 


8.9 


15.3 


.65 


1.13 


11,831 


17.15 


86.92 


493 


9.7 


13.4 


.70 


1.24 


12,002 


16.05 


84.39 


502 


9.0 


15.1 


.71 


1.23 


12,469 


14.87 


82.14 


522 


9.7 


13.3 


.63 


1.09 


12,049 


15.17 


78.12 


500 


9.5 


13.3 


.71 


1.26 


11,801 


15.75 


77.21 


470 


8.3 


15.7 


.68 


1.23 


11,769 


18.59 


84.04 


472 


8.7 


14.2 


.66 


1.27 


11,955 


16.11 


79.30 


473 


7.9 


16.1 


.62 


1.20 


12,360 


13.63 


74.59 


476 


8.8 


14.5 


.66 


1.31 


12,298 


13.62 


75.19 


480 


9.0 


14.4 


.66 


1.29 


12,423 


13.37 


75.61 


474 


9.4 


13.3 


.92 


1.18 


11,956 


17.45 


83.24 


451 


9.2 


12.5 


.76 


0.98 


11,971 


17.45 


80.99 


443 


10.0 


11.2 


.68 


1.15 


13,126 


10.24 


70.90 


487 


10.4 


12.1 



BOILERS 



133 



TABLE 28. 
RESULTS OF COAL TESTS AT ARMOUR GLUE WORKS, CHICAGO, AUG. 17, 1905. 









Cost 


Cost to 


Pounds 


Date of Test. 


Name and Kind of Coal. 


Railroad Car 
Number. 


per 
Ton 
Deliv- 


Evaporate 

1000 
Pounds of 


Water 
Evapo- 
rated per 








ered. 


Water , 


$1.00 of 
Coal. 


March 5, 1905 . . 


Williamson County 
Coal Co.'s, mine run 


C.C.C.&St.L. 
No. 26368 


$1.90 


$0.1531 


6,532 


March 3, 1905 . . 


Harden & Hafer, mine 


S. I. No. 5735 


1 70 


0.1231 


8,123 


June 14, 1905 . . 


Crerar-Clinch & Co., 
2" screenings. 


I.C. No. 88362 


1.50 


0.1293 


7,734 


June 15 1905 . . 


.do 


I. C. No. 88362 


1 50 


1218 


8,210 
8,511 


June 16, 1905 . . 


....do 


I. C. No. 88362 


1.50 


0.1175 


June 17, 1905 . . 


Brackett Coal and 
Coke Co., lump. 


C. &E. I., No. 

8891. 


1.65 


0.122 


8,197 


June 19, 1905 . . 


....do 


C. & E. I. No. 


1.65 


0.1212 


8,251 






5002 








June 20, 1905 . . 


....do 


C. & E. I. No. 

5002 


1.65 


0.1352 


7,396 


July 1, 1905. ... 


Kellyville Coal Co., 
mine run. 


C. & E. I. No. 

10030. 


1.595 


0.1355 


7,380 


July 6, 1905. . . . 


Brackett C. & C. Co., 
Keeler mine run. 


C. & E. I. No. 

12367 


1.65 


0.1236 


8,091 


July 28, 1905. . . 


Kellyville Coal Co., 
washed pea. 


C. & E. I. No. 

6211. 


1.50 


0.1285 


7,782 


July 29, 1905. . . 


....do 


C. & E. I. No. 


1.50 


0.119 


8,403 






6211 






Aug. 5, 1905 . . . 


Dering Coal Co., mine 
run. 


C. & E. I. No. 
25125 


1.575 


0.125 


8,000 


Aug. 7, 1905 . . . 


Dering Coal Co., Sulli- 
van Co., screenings. 


E. &T. H. No. 

5132. 


1.40 


0.11 


9,091 


Aug. 8, 1905 . . . 


Consolidated Indiana 
Coal Co., Sullivan 
Co., screenings. 


E. &T. H. No. 
3239 


1.35 


0.105 


9,524 


Aug. 9, 1905 . . . 


Screenings 


E &T. H. No 


1.30 


0973 


10,277 




6534 






Aug. 11, 1905 .. 


Ziegler, screenings . . . 


I.C. No. 81184 


1.50 


0.1047 


9,551 



TABLE 29. 

RELATION BETWEEN CAPACITY AND EFFICIENCY. 

(Evaporation from and at 212° F. per Square Foot of Heating Surface per Hour.) 



2 


2.5 


3 3.5 


4 5 


6 


8 


10 


12 


Probable Relative Economy, Ordinary Installation. 


100 


100 


100 


95 


90 


85 


80 


70 


60 


50 


Probable Relative Economy, Latest Improved Installation. 


95 


98 


100 


100 | 100 


99 


98 


95 


90 


85 



134 



STEAM POWER PLANT ENGINEERING 



TABLE 30. 

FLUE GAS TEMPERATURES CORRESPONDING TO FORCED CAPACITY OF BOILERS 
IN MODERN POWER PLANT INSTALLATIONS. 



Plant. 



Cambridge Steel Co 

Commonwealth Edison Co 

Detroit Edison Co 

Everett Mills 

Narragansett Electric 

Lighting Co 

National Museum 

N. Y. Central R.R., West 

Albany 

N. Y. Central &H.R.R.R 
N. Y. Edison Waterside. . . 

Old Colony St. Ry. 

Union Gas & Electric Co... 



Type of Boiler. 



B. & W. 
B. & W. 

Stirling 
Manning 

B. & W. 
Geary,W. T 

Edgemoor 
Ret. Tub. 
B. & W. 
B.& W.- 
Stirling 



Rated 

Horse 

Power per 

Unit. 



400 
650 

2365 
130 

440 

182.8 

600 
100 
650 

687.5 
542 



Heat Sur- 
face per 
Horse 
Power 

Developed 



5.14 
4.97 
4.75 
6.0 

5.5 
6.4 

5.28 

4.4 

5.48 

5.25 

4.43 



Flue 
Tempera- 
ture. 



485 
588 
651 
599 

544. 
430 

543 
630 
550 
599 

622 



Builders' 

Rating, 

Per Cent. 



194.5 

201.0 
211.3 
150.0 

180.4 
155 

193 
273 
179 
190 
227 



TABLE 31. 

PRINCIPAL DATA AND RESULTS OF TESTS ON 2365-RATED-HORSE-POWER STIRLING 

BOILERS AT THE DELRAY STATION OF THE DETROIT EDISON COMPANY. 

Tests with Roney Stoker. Resume 1 of Principal Results. 



No. of 


Length, 


Per Cent 


Test. 


Hr. 


Rating. 


1 


25 


105.0 


2 


24 


80.0 


3 


24 


113.8 


4 


30 


152.4 


5 


24 


94.0 


6 


24 


150.7 


16 


32 


98.6 


17 


16.5 


193.3 


18 


24 


195.7 


2-4 f 


90 


119.8 


5-6t 


55 


127.3 



B.t.u. in 
Coal. 



14,362 
14,225 
14,308 
13,756 
13,896 
14,037 
14,476 
14,493 
13,689 
14,098 
13,977 



Per Cent 

Ash in 
Dry Coal. 



5.98 
6.52 
7.40 
6.54 
6.89 
6.13 
9.68 
8.24 
9.81 
6.81 
6.84 



Efficiency. 



77.84 
79.88 
77.45 
75.78 
81.15 
75.28 
80.98 
76.73 
75.57 
76.13 
76.23 



Per Cent 
Steam used 
by Stoker 

Engines 

and Steam 

Jets. 



0.63 
1.58 
1.75 
1.45 
1.34 
1.39 
1.32 



Per Cent 
Combusti- 
ble in Ash. 



19.6 
17.9 
24.4 
30.8 
31.6 
26.7 
34.1 
24.6 
23.2 
25.8 
29.4 



Temp, of 

Flue Gases 

Leaving 

Boiler, 

Deg. Fahr. 



576 

480 
542 
670 
483 
662 
460 
636 
694 
572 
575 



t Including periods between tests. 

Tests with Taylor Stoker. Resume of Principal Results. 



No. of 
Test. 


Length, 
Hr. 


Per Cent 
Rating. 


B.t.u. in 
Coal. 


Per Cent 

Ash in 
Dry Coal. 


Efficiency. 


Per Cent 

Steam used 

by Stoker 

Auxiliaries.* 


Per Cent 
Combusti- 
ble in Ash. 


Temp, of 
Flue Gases 

Leaving 

Boiler, 
Deg. Fahr. 


7 


24 


151.2 


14,000 


7.03 


77.07 


2.61 


31.5 


575 


8 


24 


107.9 


13,965 


6.34 


80.28 


2.44 


27.1 


493 


9 


50 


162.8 


13,998 


6.75 


77.85 


2.87 


31.3 


574 


10 


48 


92.9 


14,188 


9.90 


77.90 


2.63 


27.2 


487 


11 


26.5 


211.3 


14,061 


9.55 


75.84 


3.41 


36.1 


651 


12 


48 


121.3 


14,010 


8.09 


79.24 


2.57 


27.6 


535 


14 


24 


185.2 


14,272 


8.71 


76.42 


2.95 


28.8 


647 


15t 


24 


123.1 


14,213 


8.34 


74.90 


2.77 


30.1 


561 


7-9 1 


109 


140.0 


13,983 


7.22 


77.66 


2.68 


29.9 


545 


10-llt 


80.5 


132.8 


14,095 


9.58 


75.66 


3.04 


31.1 


542 



* Engines driving stokers and steam-turbine driving fan. 

t In test No. 15 the fires were banked for 7j hours and the averages include this period. 

% Including periods between tests. 



BOILERS 



135 



In nearly all stations the boilers must have sufficient overload capac- 
ity to take care of peak loads or to allow some of the boilers to be shut 
down for cleaning or repairs, since the installation of sufficient rated 
boiler capacity would be expensive and in many instances prohibitive in 
cost. In small stations, however, too large a boiler capacity frequently 
is to be preferred to an overloaded installation, since the extra first cost 
of the former may be less than the loss due to poor efficiency and depre- 
ciation in the latter. 

As far as forcing is concerned the fire-tube boiler is as effective as 
the water-tube, more depending upon the furnace, grate surface, draft 
and character of fuel than upon the type of boiler. All boilers are 
subject to more or less priming at heavy overloads, and the overload 
capacity is often limited on this account. 

The Forcing Capacity of Fire-Tube Boilers: F. W. Dean, Trans. A.S.M.E., 26-92. 

Increasing Capacity of Steam Boilers: Kreisinger and Ray, Power, May 24, 1910. 



•+3 








































«. 


650 


Q 






600 B & W BOILER EQUIPPED 

WITH L TYPE GREEN CHAIN 

GRATE AT THE INTERBOROUGH 

RAPID TRANSIT CO. ,.N.Y. 

MARCH 1912 




















& 

P- 


















/ 






600 


O 
















/ 








3 




























V 


/ 










550 


§80 

u 
o, 75 




























/ 


/ 












1 










4- 












V 


/ 
















.500 


o 
H70 

V, 




h 


oiler 


and 


Furn 


ace j 


Cffici 


sncy 




1 


ty 


f 












| 






















4? 






1 




















& " 

a 












^V»g 


<sg2 






























S 
"3 

SB 
































Na 


tura 


lDr 


aft 
















































3 


00 


4 


JO 


5( 


)0 


6( 


X) 


7( 


)0 


8( 


30 


9( 


X) 


1,( 


XX) 


1,1 


00 


u 


00 


M 


00 



Capacity, Boiler Horse-power 
Fig. 57. Effect of Rate of Driving on Capacity. 

82. Thickness of Fire. — For a given furnace and boiler, quality and 
size of fuel and intensity of draft, a certain depth of fuel will give maxi- 
mum efficiency. Too thin a fire results in an excess of air and too thick 
a fire in a deficiency, the economy being lowered in either case. On 
account of the number of conditions upon which the proper thickness 
depends, it can only be determined for a particular case by actual test, 
the available data being insufficient for drawing conclusions. The 
curves in Fig. 63 are plotted from a series of tests made on a 350-horse- 
power Stirling boiler equipped with chain grate at the power plant of 
the Armour Institute of Technology. The damper was left wide open 
throughout the test and the speed of the grate kept constant. Ratio 
of grate to heating surface, 1 to 42. Carterville washed coal No. 4 was 



136 



STEAM POWER PLANT ENGINEERING 



75 


















































i 


i 

• • 
• 






• 
• 


• 


• 




70 


















• 


• 


• 




•• 


• • 
















• 




• * 


• 


.•x 


• .^ 






• 


• 




a 

O 
<3 65 






• 




• 


• 
• 


• 


••• 


• 


< 

• • 


























• 


S» 


• 
• 


• 


• 


• 


• 
• 










a 
6 60 






• 






f • 


• • 
• 
•• 


• 
• 
< 


• 
i 


• 


• 




























• 


















s 

s 
Ph 55 








• / 










500 H.P. Ib. & W..Boiler 
5000 Sq. Ft. Water Heating Surface 
940 -, »» Superheating 
Standard Chain Grate and Setting 
90 Sq. Ft. Grate Surface 




■a 

a 

eS 


















o 
«50 
















Various Grades ot <Joal 
Fisk Street Station, Common Wealth-Edison Co. 
Chicago 111. 


































45 


•y 


_ 




• 



























































400 



Fig. 58. 



500 



600 



700 800 000 

Boiler Horse Power 



1000 



1100 



1200 



Relation Between Efficiency and Capacity; 500 H.P. Boiler, Fisk Street 
Station, Commonwealth Edison Co., Chicago. 



■a u.o 



c 




r> 


10.8 


n 




*o 


I0.fi 






i-i 




I 


L0.4 












0,2 


a 








£ 


10 


a 




K 




«4-I 


9H 


a 












2 


9.6 


5. 








H 


9.4 






a 




"3 


9.2 

















60C 


H.P 


B.&W. 


BOILER 














h 










6008 C ; 0£ 

EQUIPPED WITH RONEY STOKERS 

AT THE 59th ST STATION 

OF THE INTERBOROUGH RAPID 

TRANSIT CO., N.Y. 












' 


















i 










































• \ 






























1 


sir 


IGLE 


STO 


<ER 


















^ \ 


^1 








J 




100 


5q.Ft.Gra1 


e Surface 


















*\ 










Coa 


1 14250 B.T 


.U.per Po 


and 










DOUBLE 
Front, 


STOKER N 
100 Sq.Ft. 










Nat 


ural 


Draft 
















Re 


ar, 


80 a 


q.Ft. 




X. 







































w 600 680 760 840 920 1000 1080 1160 1240 1320 

- Boiler Horse Power 

Fig. 59. Effect of Rate of Driving on the Efficiency of a 600 H.P. B. & W. Boiler. 



BOILERS 



137 



used in all tests. The curves in Fig. 64 refer to the performance of a 
150-horse-power water-tube boiler equipped with chain grate at the 
University of Illinois Engineering Experiment Station at Urbana. 



1300 



1200 



1100 



900 



800 



BOO 



600 











































/ 
































A 


/ 

* 








b/ 






C 


















7 










> 
























/ 


D 


DUBL 
F 
TJ, 


EST 
ront, 
3ar._ 


DKEF 
100 Sc 
80 So 


/ 


/ 


• 








^ O 




*^<> 




Astatic Draft over Eire 
B u M in Rear Pass 
C Velocity Draft in Flue 

Immediately outside of 

"Boiler Setting 














yT < 


■ ^s 
















B/ 


























V 








u 


/ 




























SINS' 


'jy£\ 


"OKE 
100 S 


q.Ft. 
























o 




/ 


> 




























// 


/ 








600 H.P. B.& W- BOILER* 

6008 SQ.FT. HEATING SURFACE 

RON EY STOKERS 

59th ST. STATION,,! NTERBOROUGH 

RAPID TRANSIT CO., N-Y- 








/ 


<y y 


















oo 





























0.1 



0.2 



0.1 



0.B 



0.3 0.4 0.5 

Draft in Inches of Water 

Fig. 60. Influence of Draft on the Capacity of a 600-Horse-power B. & W. Boiler. 

The curves in Fig. 65 are plotted from a series of tests on a 500-horse- 
power Babcock and Wilcox boiler equipped with chain grate at the Fisk 

&5 



75 



| 70 



65 



Fig. 61. 



400 500 600 700 800 900 1000 1100 1200 
Boiler Horse power 

Relation between Efficiency and Capacity — Oil Fuel. 



Street station of the Commonwealth Edison Company, Chicago, 111. 
In these tests the conditions of operation are not exactly comparable, 
but they serve to show the variation of economy with thickness of fire 



138 



STEAM POWER PLANT ENGINEERING 



in each case. In general, with natural draft, fine sizes of coal necessi- 
tate thin fires, since they pack so closely as to greatly restrict the draft. 
Thin fires require closer attention to prevent holes being burned in 
spots, and respond less readily to sudden demands for steam, but have 



'80 



o 



70 



°-65 



o — -^^^ 

o 



Fig. 62. 



3 4 5 6 7 

Water Evaporated per sq, ft. of Heating 
Surface from and at 212° Fahrenheit 
Relation between Efficiency and Evaporation — Oil Fuel . 



the advantage of letting the air required pass through the grate, whereas 
thick fires often require air to be supplied above the grate to insure 
complete combustion. Thick fires require less attention and hence are 
preferred by firemen. Where sufficient draft is available thick fires 



6 

H 

1 a 50 

■2^-40 



Fig. 63. 



















i 






































e 




Cat 


>acj 


"tF 


I — j 


> 










































































































































1 


l 












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1 




































\( 


1 
















"o 


: 


Sffi 


3ier 


icy 


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— C 


> 


















X 


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c 


3 







































































































































100 gj 

o 



4 5 

Thickness of Eire, Inches 



Effect of Thickness of Fire on the Capacity and Efficiency of a 350-Horse- 
power Stirling Boiler, Equipped with Chain Grate. 

are more efficient than thin ones, as the air excess is more readily con- 
trolled. 

83. Influence of Initial Temperature on Efficiency. — In general the 
higher the initial temperature of the furnace the greater will be the 
efficiency of the heating surface, since the heat transmitted varies almost 
directly with the difference of temperature between the water and the 
products of combustion. If the heating surface is properly distributed 






BOILERS 



139 



so that the final temperature of the escaping gas remains constant, the 
efficiency of the boiler and furnace will increase as the initial temper- 
ature increases, though not in direct proportion. This is on the assump- 
tion that the amount of heat generated per hour is the same through- 
out all ranges in temperatures. With a condition where the amount of 
heat generated remains constant and the initial temperature varies, 
the final temperature of the escaping gases remains practically constant, 



Horse- Power Developed 
(34 M Lbs. of Water Evap. into 
Dry Steam F. & A. 212° per Hr.=l H.P.) 

o o o o o o 




















O ( 


y^i 


y 
























OS 










Bate 


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city oi 


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7.5 

o ij tj 
^ S3 O 

> <* £h 

H $ © 6.5 
£ Z * 

2 fi £ 6.0 
I'll 

0.0 








o 
























o ^^ 
























• 










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2£e_ 


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• 


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6-in. 


Fire 


• 


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v 


A 


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15 20 25 • 30 

Dry Coal per Sq. Ft. of Grate Surface per Hr.-Xbs. 



35 



40 



Fig. 64. 



Effect of Thickness of Fire on the Capacity and Efficiency of a 150-Horse- 
power Water-tube Boiler. 



and in such cases high initial temperatures are productive of high 
boiler and furnace efficiencies. In practice these conditions are seldom 
realized and high furnace temperatures are not necessarily productive 
of high boiler and furnace efficiencies. Some tests show a decided gain 
in efficiency with the higher furnace temperatures ("Some Perform- 
ances of Boilers and Chain-grate Stokers, with Suggestions for Improve- 
ments, " A. Bement, Jour. West. Soc. Engrs., February, 1904), and 
others show little if any improvement ("A Review of the United States 
Geological Survey Fuel Tests under Steam Boilers," L. P. Brecken- 



140 



STEAM POWER PLANT ENGINEERING 



ridge, Jour. West. Soc. Engrs., June, 1907). The majority of high- 
efficiency records, however, are associated with high furnace temper- 
atures. 





















































































1000 




































































































































































































































































































































» 800 


































































































































































£ 


















































































I 

§ 600 


































































































































































































































































































































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' 


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•"s. 


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400 












V S 
















" 


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K. 


























































































































200 















































































































































































































































































































































































































































































































































































































































































































































































































































t» 80 


































































































































































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3 








































































































































































































40 








































































































































A Boiler 14 Tubes High 
B « 9 «• 
















































































































































































J 


DU 


\ 


w 


B. 


K. 


11-539 





7 8 9 10 

Thickness of Tire in Inches 



12 



Fig. 65. Effect of Thickness of Fire on the Capacity and Efficiency of a 500-Horse- 
power Babcock and Wilcox Boiler. 



TABLE 32. 

TEMPERATURE DROP OF GASES THROUGHOUT BOILER. 
(650 H.P., B. & W. Boiler, Waterside Station of N. Y. Edison Co.) 





Boiler 
and Grate 
Efficiency. 


Temperatures, Degrees F. 


Per Cent 

of 
Rating. 


Furnace 
Temper- 
ature. 


Middle 
First 
Pass. 


Top 

First 
Pass. 


Top 

Second 
Pass. 


Middle 

Second 

Pass. 


Middle 
Third 
Pass. 


Flue. 


117.3 
126.7 


78.5 
79.6 
79.8 
77.1 
75.3 
77.6 
76.5 
72.7 


2336 


866 
893 
888 
889 
913 
956 
939 
1051 


655 
689 
681 
682 
694 
723 
700 
751 


619 
646 
633 
642 
655 
660 
634 
700 


511 
526 
521 
519 
512 
547 
523 
578 


471 
485 
481 
479 
486 
512 
493 
541 


455 

473 


128.6 
131.0 
131.0 
137.4 


2420 
2455 
2430 


468 
468 
473 
492 


142.2 




475 


185.3 


2530 


519 



BOILERS 141 

84. Cost of Boilers and Settings. — Figures giving the cost of boilers, 
irrespective of type, at so much per horse power are misleading, since 
the cost does not increase in the same ratio as the power. The wide 
variation in cost on the horse-power basis is partly due to the difference 
in rating. For instance, Scotch-marine boilers are ordinarily rated at 
8 square feet of heating surface per horse power and return tubular 
boilers- at 12 square feet. The price approximates one dollar per 
square foot of heating surface for all boilers over 100 horse power. 
The cost of water-tube and fire-tube boilers may be roughly estimated 
by the following formulas (C. H. Benjamin, Engr. U. S., Nov. 15, 
1902): 

(A) Cost in dollars = 500 -f- 9.2 X rated horse power. 

(B) Cost in dollars = 500 + 8.5 X rated horse power. 

(C) Cost in dollars = 100 + 6.5 X rated horse power. 

(D) Cost in dollars = 100 + 5.0 X rated horse power. 

(A) Horizontal water-tube boilers, 125 pounds pressure, 10 square feet heating 
surface per horse power. 

(B) Vertical water-tube boilers, other conditions as in (A). 

(C) Horizontal return tubular boilers, 12 square feet heating surface per horse 
power. 

(D) Small vertical fire-tube boilers. 

The cost of Scotch-marine boilers rated on a basis of 8 square feet 
per horse power may be estimated by means of formula (A) . 

The cost of plain settings may be roughly approximated as follows: 
Horizontal water tube. 

Cost = 400 + 0.8 X rated horse power. 
Return tubular: 

Cost = 300 + 0.7 X rated horse power. 

85. Selection of Type. — Boilers constructed by builders of good 
repute are usually designed for safety, durability, and capacity, and 
rigid specifications and inspection of material and workmanship are 
ordinarily not necessary, as the makers' reputations are sufficient 
guarantee of their worth. Marked departure from standard designs 
must necessarily be specified, but in most cases instructions are limited 
to the extent of heating and grate surface, the character of the furnace, 
and arrangement of setting. Numerous tests on various types of boilers 
show practically the same efficiency provided the furnaces and boilers 
are properly designed, so that the relative merits may be considered 
with reference to (1) durability; (2) accessibility for repairs; (3) facil- 
ity for cleaning and inspection; (4) space requirements; (5) adapta- 
bility to the type of furnace and stoker desired; (6) capacity; and 
(7) cost of boiler and setting. For high pressure, 150 pounds per square 



142 STEAM POWER PLANT ENGINEERING 

inch or more, the water-tube or some form of internally fired boiler in 
which the shell plates are not exposed to the high temperature of the 
furnace is considered safer than the horizontal tubular boiler because 
the shell plates and the seams of the latter must be of considerable thick- 
ness in large units, and being exposed to the hottest part of the fire are 
likely to give trouble, especially if the water contains scale or sediment- 
forming elements. Return tubular and stationary locomotive boilers 
are seldom made in sizes over 200 horse power and hence are not to be 
considered for large units. For sizes over 150 horse power where over- 
head room is limited the return tubular boiler is most commonly in- 
stalled, unless high pressure is essential, in which case the internally 
fired Scotch-marine boiler is peculiarly adaptable. The water-tube 
boiler is usually employed in large central stations for high-pressure units 
of 300 to 2500 horse power. The, particular type of water-tube boiler 
is to some extent a matter of personal taste on the part of the engineer. 
For small powers and for intermittent operation, small vertical or 
horizontal fire-box boilers have the advantage of low first cost. The 
small air leakage and radiation losses give internally fired boilers an 
advantage over externally fired fire-tube or water-tube types, but this 
is partly offset by the greater extent of regenerative surface in the setting 
of the latter.* Internally fired boilers are more expensive than the 
externally fired, though the extra cost of setting and foundation in the 
latter may bring the total cost of the entire equipment to practically 
the same figure. The design and installation of the boilers and furnaces 
should be left at the outset to a capable engineer. 

Makers usually request the following information from intending pur- 
chasers : 

1. Steam pressure desired. 

2. The quantity of steam demanded. 

3. The kind of fuel to be burned. 

4. The type of furnace or stoker. 

5. The nature and intensity of draft. 

6. Nature of setting. 

7. Probable temperature of the feed water. 

The complete specifications for a return tubular boiler are given in 
Chapter XIX. 

86. Grates. — Grates may be divided into three general classes, 
namely, stationary, rocking, and traveling grates. The latter are 
treated in Chapter IV. Stationary grates are generally made of cast- 
iron sections in a variety of shapes as illustrated in Fig. 66. The bars 

* At the power plant of the Cosmopolitan Electric Co., Chicago, the brick settings 
of the boilers (500 h.p. B. & W.) are completely incased with riveted boiler plates. 



BOILERS 



143 



are ordinarily from 3 inches to 4 inches deep at the center (this makes 
them strong enough to carry the load caused by the weight of the fuel 
without sagging even when the top is red hot), f inch wide at the top, 
and taper to f inch at the bottom to enable the ashes to drop clear. 

TABLE 33. 
AIR SPACES AND THICKNESS OF GRATE BARS. 



Size and Kind of Coal. 



Screenings 

Anthracite — 

Average 

Buckwheat .... 

Pea or nut 

Stove 

Egg 

Broken 

Lump 

Bituminous, average 
Wood — 

Slabs 

Sawdust 

Shavings 



Width of Air 


Thickness of 


Spaces. 


Grate Bars. 


(Inch) 


(Inch) 


1 


1 


\ 


1 


1 


I 


* 


h 


f 


i 


I 


h 


I 


i 


1 


I 


1 


! 


f 


f 


Itof 


I 


*tof 


1 



The width of the air space is determined by the size of the fuel to be 
used, the average proportions being given in Table 33. 

The "Tupper" and " Herringbone " grate bars are stiff er and less 
likely to warp than the common form, but are not so readily sliced and 



1 1 1 1 






COMMON BAR 


WffffffffffffA 


[«««««««« [ 




TUPPER HERRINGB 


ONE 


.; / / 


• • •• 


v.v/.v.v.v.v.v.v.v.v C 


•••••••••••••••••••• \ V 


! 1 


SAW 
Fig. 66. Types 


_1 L 

-DUST 

of Grate Bars. 


i i 



therefore not so convenient with coal that clinkers badly. Sawdust 
or pinhole grates are used in burning sawdust, tanbark, and very 
small sizes of coal. Grates are often set horizontally and the bars are 
held in place simply by their own weight, but long grates are best placed 



144 



STEAM POWER PLANT ENGINEERING 



sloping toward the rear to facilitate firing. The front of the grate when 
designed for bituminous coal is often made solid, this portion being 
called the "dead plate." It serves to hold the green fuel until the 
hydrocarbons have been distilled off, when the charge is pushed back 
on the open grate at the time of next firing. The length of a single bar 
or casting should not exceed three feet. The length of grate may be 
made of two or three bars and should not exceed 6 feet with bituminous 
coal, as this is the greatest length of fire that can be readily worked by 
a stoker. With buckwheat anthracite furnaces 12 feet in depth are not 
unusual, as anthracite fires require no slicing. 

The disadvantage of using stationary grates is that the fire is not 
easily cleaned. Unless the air spaces are kept free of clinkers and 
ashes, combustion is hindered and the fire rendered sluggish. Frequent 
cleaning, however, is wasteful of fuel and reduces the furnace efficiency 
by letting in a large excess of air every time the fire door is opened. 




1 1 v v l v '<<<<<< /=?=* 



Fig. 67. A Typical Shaking Grate. 

87. Shaking Grates. — Shaking grates have the advantage of per- 
mitting stoking without opening the fire door and require less manual 
labor than stationary grates. There is a great variety of sectional 
shaking grates on the market and some of them are made self-dumping. 
One of the best-known types is illustrated in Fig. 67. Each row or 
section of grate bars is divided into a front and a rear series by twin 
stub levers and connecting rods. An operating handle is adapted to 
manipulate either one or both of the levers in such a manner that the 
front and rear series may operate separately or together. The shaking 
movement causes no increase in the size of the openings and hence pre- 
vents the waste of fine fuel. Ordinarily the width of the grate is made 
equal to two or more rows of grate bars so that the live fire may be 
shoved sidewise from one row to the other when cleaning. A depth of 



BOILERS 



145 



fire of from 6 to 10 inches is carried according to the nature of the fuel 
and the available draft. 

Grate Bars: Engr. U.S., Nov. 1, 1906, p. 728, Jan. 1, 1907, p. 68; Am. Elecn., Jan. 
1904, p. 269. 

88. Blow-Offs. — Boilers must be provided with blow-off pipes for 
draining off the water and for discharging sediment and scale-forming 
material. The " bottom blow" is ordinarily an extra-heavy pipe of 





Fig. 68. Horizontal Blow-off 
Connection to Head. 



Fig. 69. Vertical Blow-off 
Connection to Shell. 



suitable diameter connected to the lowest part of the boiler and fitted 
with a valve or cock, or both. (See Fig. 512.) Fig. 68 shows an 
arrangement of horizontal blow-off connected to the head of a return 












1 

1 ; 


m 


I 













BLOW OFF 



Fig. 



70. Blow-off Connection with 
Circulating Pipes. 



Fig. 71. Blow-off Tank and 
Connections. 



tubular boiler and Fig. 69 a vertical blow-off connected to the shell. 
The latter is the better arrangement. The blow-off pipe where it 
passes through the back connection is covered with magnesia, asbestos, 



146 



STEAM POWER PLANT ENGINEERING 



or fire brick. When exposed to the action of extremely hot furnace 
gases as in forced-draft installations, the arrangement illustrated in 
Fig. 70 is sometimes used to prevent the pipe from burning out. When 



Water Level —_-^ 




Fig. 72. Surface Blow-Off. 

the blow-off cock is shut and the valve on the vertical branch is open 
there is a continuous circulation of water. Where boilers are arranged 
in batteries, the battery may have a common outlet for the blow-off 
pipes, as illustrated in Fig. 512. Usually the blow-off pipes may dis- 
charge into the open air, but this is not permissible in large cities, nor 
is it lawful to blow directly into the sewer. In this case the water and 
sediment may be discharged into a blow-off tank and permitted to cool 
before delivery to the sewer, as illustrated in Fig. 71. 




Fig. 73. Buckeye Skimmer. 

"Surface blows" are often installed to remove scum, grease, and float- 
ing or suspended particles of dirt. The bell-mouthed shape shown in 
Fig. 72 permits the skimmer to accommodate itself to varying water 
level, and it is sometimes provided with a float and with a flexible joint, 
Fig. 89. 

89. Damper Regulators. — For maximum furnace efficiency the draft 
must be regulated to burn just enough fuel to supply the steam 



BOILERS 



147 



required. Where forced draft is employed this is done by regulating 
the speed of the blower. With natural draft it is the usual practice to 
regulate the draft by means of dampers placed in the uptake, and in 
order that the regulation may be effective it should be automatic. 
Automatic dampers are economical and useful and are particularly 
desirable in plants where the demand for steam fluctuates rapidly. 
There are several successful types on the market, some operated by 



OR 


u __j 1 

■ N 


~D 


Steam n^" D _£__ 


nip 


_L 


I Jill "' A 

a 


Exhaust \D If f 1 " ™ ~j 
Water Supply . , . ^ 
— on°D 


-W— 



Fig. 74. Kitts Hydraulic Damper Regulator. 

water pressure and others by direct boiler pressure. Fig. 74 illus- 
trates such a mechanism. Full boiler pressure acting at all times on a 
diaphragm A raises or lowers a weight W attached to arm D according 
as the steam pressure increases or decreases. Arm D actuates a small 
valve V which controls the supply of water to chamber B. A diaphragm 
in chamber B raises and lowers the damper as the water pressure varies, 
a drop of 0.5 pound being sufficient to open the damper to its maximum. 
The steam diaphragm has a movement of only 0.01 inch and the water 
diaphragm 0.5 inch. When properly adjusted and given proper atten- 
tion automatic dampers work in a very satisfactory manner. 

Fig. 75 shows a section through the Tilden damper regulator, illus- 
trating the principles of the steam-actuated type. The device is con- 
nected directly to the boiler by pipe A. The pressure on piston B is 
balanced by spring C under normal conditions of operation. Any 
variation from the normal will cause the rod R to move up or down, so 
that the dampers are opened or closed in proportion to the change in 
pressure. The chamber N is separate from chamber M, so that steam 
or water cannot come in contact with the spring. Piston D acts as a 
guide only. In a recent design of this device the regulator is hydrauli- 
cally actuated and simultaneously operates the damper and the stoker 
engine thereby automatically proportioning the air and fuel supply to 
the load requirements. 

90. Water Gauge. — The water level in a boiler is usually indicated 
either by a gauge glass, by try cocks, or both, connected directly to the 



148 



STEAM POWER PLANT ENGINEERING 



boiler as in Fig. 1, or to a water column or combination as in Fig. 76. 
Each gauge-glass connection should be fitted with a stop valve which 
may be closed in case the tube breaks. In large boilers these valves, 
usually of the quick-closing type, are conveniently operated from the 
boiler-room level by means of a chain attached to the valve stem. 
Self-closing automatic valves are frequently employed, one type being 
illustrated in Fig. 77. If the glass breaks the outrush of steam forces 




N 




Fig. 75. Tilden Steam-actuated 
Damper Regulator. 



Fig. 76. Simple Water Column. 



the ball against a conical seat and shuts off the supply. When a new 
glass is inserted the ball is forced back by slowly screwing in the valve 
stem. Hinged valves mechanically operated from without are con- 
sidered more reliable than ball valves, as scale is less likely to render 
them inoperative. 

Try cocks or gauge cocks are set at points above and below the desired 
water level, preferably connected directly to the boiler shell, but some- 
times to a water column as in Fig. 76. The water level is ascertained 
by opening the cocks in succession. 



BOILERS 



149 



The objection to the latter arrangement is that accident to or a 
stoppage of the piping renders both gauge glass and try cocks useless. 
Water columns should be blown out once a day, and the gauge cocks 
opened to see that the height of the water indicated tallies with that 
shown by the glass. Some engineers prefer two separate columns to 
each boiler and no cocks, others rely solely upon cocks. 




r\ 



^> 



Fig. 77. Water Gauge with Self-closing 
Valve. 




v WATER 

Fig. 78. Combined Water Column 
and High- and Low-water Alarm. 



The water column shown in Fig. 78 has an alarm whistle, controlled by 
two floats, which gives a high- and low-water alarm. Numerous devices 
of this class are on the market but they are usually regarded as unre- 
liable and most engineers are content to depend upon water gauge and 
try cocks. 

Water Gauges and Columns: Mach., Sept., 1905, p. 31; Power, Aug., 1905, p. 483; 
Am. Elecn., July, 1904, p. 359; Engr. U. S., Jan. 1, 1907, p. 58. 

91. Fusible or Safety Plugs. — Fusible or safety plugs as illustrated 
in Fig. 79 are brass plugs provided with a fusible metal core. They 

Inside-T-y-pes s / ©utside-Eypes s 



J i ' : i'V"-'-:4 






\ \ } r' I 




Fig. 79. Types of Fusible Plugs. 



are inserted in the shell or tubes at the lowest permissible water line. 
When covered by water the heat is conducted away sufficiently fast to 
keep the temperature below the fusing point, but when uncovered the 
low conductivity of the steam prevents the rapid withdrawal of heat, 



150 



STEAM POWER PLANT ENGINEERING 



whereupon the alloy melts and the blast of escaping steam gives warn- 
ing. The melting point of fusible metals being sometimes uncertain, 
plugs occasionally blow out without apparent cause and at other times 
fail to act when shell is overheated. Fusible plugs are required by law 
in many cities. 

92. Mechanical Tube Cleaners. — Although purifying plants, boiler 
compounds, and the like are preventive of scale formation to a great 
extent, experience shows that the most satisfactory method is to use 
mechanical tube cleaners for cutting or breaking the scale. The prin- 
ciples of construction of these devices vary widely according to the 
types of boilers in which they are used, and depend upon the nature of 




C D F U 

Fig. 80. Mechanical Tube Cleaner — Hammer Type. 

the duty which they must perform. They may be conveniently divided 
into two classes: 

1. Those which loosen the scale by a series of rapid hammer blows, 
Fig. 80. 

2. Those which cut out the scale by a revolving tool, Fig. 81. 

The hammer device is applicable to either the water- or fire-tube type 
of boiler, but the revolving cutter is applicable to the water-tube only. 
Steam, compressed air, or water under pressure may be used as the 
motive power, though the latter is the most convenient and satisfactory. 

Referring to Fig. 80, the hammer head J is given a rapid motion, 
which may reach 1500 vibrations per minute, and subjects the tube 
to repeated shocks, thereby cracking the brittle scale and jarring it 
loose from the water surface of the tube. The cleaner head is attached 
to a flexible pipe of sufficient length to enable it to be pushed from one 
end to the other. Even if carefully manipulated the hammer is apt to 
injure the tube by swaging it to a larger diameter, producing crystalliza- 
tion in the metal and causing leaks where the tubes are expanded into 
the sheets, hence its use is not to be recommended. 

Hydraulic turbine cutters are made in many designs, one of which 
is shown in Fig. 81. The cylindrical casing D contains a hydraulic 
turbine consisting of a fixed guide plate which directs the water at the 



BOILERS 



151 



proper angles upon the vanes of the turbine wheel T. The cutters C 
revolve at high speed and chip the scale into small pieces. The stream 
of water flowing from the turbine envelops the cutters, keeps their 
edges cool, and washes away the scale as fast as it is detached. Differ- 
ent styles of cutter wheels are furnished with each cleaner so as to 
adapt the device to all kinds of scale formations. In well-managed 
plants scale is not permitted to deposit to a thickness greater than T \ 
to T ^ of an inch. 




*ty*Y»t»{», 



ik 



Fig. 81. Mechanical Tube Cleaner — Turbine Type 



The soot and cinders which accumulate on the inside surface of fire- 
tube boilers are removed by mechanical scrapers, brushes, or steam-jet 
blowers. (For a description of these devices see Power, Dec. 6, 1910, 
p. 452.) 

The tubes of a water-tube boiler are cleaned externally by means of 
a steam jet. 

Arches, Firebrick Furnaces: Power, Feb. 20, 1912. 

Blow-off Connections: Locomotive, Oct., 1906; Elec. World, Nov. 2, 1907; Nat. 
Engr., June, 1904; Eng. Rec, May 9, 1908; Steam, Feb., 1911. 

Bracing: Boiler Maker, Feb., 1912; Mach., Sept., 1903, p. 18, Oct., 1903, p. 83; 
Power, Jan., 1903, p. 24, Oct., 1905, p. 611, Nov., p. 687; Eng. News, Dec. 15, 1904, 
p. 533; Engr. U. S., Jan. 1, 1907, p. 18; Prac. Engr., Jan., 1907. 

Boiler Cleaning: Am. Elecn., Dec, 1900, April, 1904, p. 174; Power, May and 
Oct., 1905, Aug., 1906, p. 465; Locomotive, Oct., 1904; Boiler Maker, Aug., 1905; 
Engr. U. S., Jan. 1, 1907, p. 109; Power & Engr., Nov. 8, 1910, p. 1975. 

Boiler Design: Engr. U. S., Jan. 15, 1902, p. 59; Eng. Mag., May, 1904, p. 233; 
Eng. Rec, July 14, 1900, May 18, 1901, p. 467, Oct. 12, 1901, p. 347; Power, Oct., 
1901, p. 14, March, 1906, p. 147; Am. Mach., April 21, 1904. 

Boiler Dimensions: All Types of Stationary Boilers: Eng. U. S., Jan. 1, 1907, p. 10, 
Aug. 1, 15, 22, 1903. Small Marine Boilers: Am. Mach., Sept. 3, 1896, p. 823. 
Tubular Boilers: Mach., Oct., 1902, p. 94. 

Circulation in Boilers: Eng. Rec, July 20, 1901; Cassier's Mag., Jan., 1905; Elec 
Rev., Lond., April 4, 1902; Engng., April 18, 1902; Engr., Lond., Nov. 6, 1903; 
Am. Mach., Jan. 14, 1897, p. 40, Sept. 20, 1900, p. 910; Eng.. News, Jan. 18, 1900, 
p. 40; Trans. A.S.M.E., 7-814, 9-489; Engr. U. S., Oct. 15, 1907. 

Damper Regulators: Engr. U. S., Jan. 1, 1907, p. 58; Elec. Wld., May 2, 1908. 

Domes: Engr. U. S., Jan. 1, 1907, p. 27. 



152 STEAM POWER PLANT ENGINEERING 

Classification of Boilers and Comparison of Types: Engr. U. S., Jan. 1, 1907; 
Min. Rept., Feb. 21, 1907. 

Explosions, Cause of: Power, Mar. 26, 1912; Boiler Maker, July, 1910. 

Furnace and Settings: Power & Engr., Jan. 3, 1911, p. 2; Elec. World, Sept. 7, 
1907; Am. Elecn., Jan., 1902, p. 10, Nov., 1903, p. 557, July, 1904, p. 339; Engr. 
U. S., July 15, 1905, p. 471, Sept. 15, 1905, p. 622, Aug. 1, 1906, p. 491, May 15, 
1906, Jan. 1, 1907; Power, March 24, 1908, p. 445, June, 1905. 

Inspection: Power, Jan., 1906, p. 32; Engr. U. S., Feb. 15, 1907; Trans. A.S.M.E., 
4-142; Boiler Maker, Feb., 1911. 

Riveted Joints: Power, March, 1906, p. 147, April, 1906, p. 227; Engr. U. S., 
Jan. 1, 1907, p. 21, Aug. 15, 1907, p. 784; Boiler Maker, June, 1906, Dec, 1907; 
Prac. Engr., Dec. 13, 1907. 

Safety Valves: See paragraph 390. 

Specifications: Power, Dec, 1905, p. 728; Nat. Engr., May 15, 1903, p. 367; 
Boiler Maker, Sept., 1906, p. 243. 

Thickness of Boiler Plate: Am. Mach., Jan. 16 and Feb. 27, 1902; Trans. A.S.M.E., 
22-127, 15-629, 24-921; Eng. News, Jan. 31, 1901, p. 121. 

Testing: See A.S.M.E. Code for conducting Standard Boiler Trials, reprinted in 
Appendix B. See also Power and Engr., Feb. 23, 1909. 

Testing to Destruction: Eng. News, Feb. 1, 1912. 

Bridge Walls in Theory and Practice: Power and Engr., Mar. 9, 1909, p. 452. 

Soot Blowers and Soot Suckers: Power and Engr., Dec. 6, 1910, p. 2143. 

Installing Tubes in Boilers: Power and Engr., Nov. 1, 1910. 



CHAPTER IV. 

SMOKE PREVENTION, FURNACES, STOKERS. 

93. Current practice shows that bituminous coals high in volatile 
matter can be efficiently burned without smoke, provided the furnaces 
are properly designed, and the necessary attention is given draft re- 
quirements and stoking. 

The problem of smoke abatement is a comparatively simple one for 
large plants equipped with mechanical stokers and provided with 
ample draft, even for widely fluctuating loads, but for hand-fired plants 
it depends largely upon skillful manipulation by interested and efficient 
firemen. The order of intelligence demanded for this work is out of 
all proportion to the wages paid. In many small plants — and these 
are usually the most obstinate smoke offenders — the fireman handles 
as much as a ton of coal per hour by hand, besides caring for the feed 
pumps and water levels, keeping the boilers clean, and removing the 
ash. The boiler room is frequently poorly lighted and poorly venti- 
lated. It is, therefore, not surprising that the fireman seldom worries 
about the smoke problem. A better wage scale and more consider- 
ation for the fireman might do a great deal toward abating the smoke 
nuisance. 

Since the loss in heat due to visible smoke is usually less than one 
per cent, and seldom greater than two per cent of the heat value of the 
coal, it is a common statement among owners of power plants that it 
is cheaper to smoke than to operate without smoke. This is undoubt- 
edly true in many cases where smokeless combustion can be secured 
only by admitting a considerable excess of air with a consequent loss 
in economy frequently greater than that due to incomplete combustion 
and smoke, but if proper attention is given to the various factors in- 
volved practice shows that smokeless combustion can be effected with 
high boiler and furnace efficiency. 

In order that combustion may be smokeless and efficient, the volatile 
gases and separated free carbon must be brought into intimate contact 
with the proper quantity of air and maintained at a temperature above the 
ignition point until oxidation is complete before they are brought into con- 
tact with the heat-absorbing surfaces of the boiler. Mere excess of air 
will not effect smokeless combustion, even if the gases and air are thor- 
oughly mixed, if the temperature is prematurely reduced below that neces- 

153 



154 



STEAM POWER PLANT ENGINEERING 



sary for combustion by contact with the heat-absorbing surfaces of the 
boiler. 

Smoke may be produced, therefore, by 

1. An insufficient amount of air for the perfect combustion of the 
volatile gases. This is primarily a function of the draft. 

2. An imperfect mixture of air and combustible. 

3. A temperature too low to permit complete oxidation of the volatile 
combustible. 

The curves ia Fig. 82 give an idea of the distribution of smoke 
production by various industries in Chicago, in 1910. Rigid enforc- 
ment of the smoke ordinance has reduced the nuisance to a consider- 
able extent, so that the distribution at this date differs somewhat from 
that shown in the curves. 



1 Central 
"District , 



-2 Miscellaneous 
-•Power Plants 



3 Flats 

4 Domestic 

5 Special Furnaces 



6 Railroads 

7 Boats 




Fig. 82. Smoke Distribution in Chicago Plants. 



Smoke-preventing devices may be divided into two classes: (1) those 
which may be conveniently attached to plants already in operation 
without material modification of the furnace, such as steam jets and 
other means of mixing the air and combustible gases, admission of air 
through the bridge or side wall, and mechanical draft; and (2) those 
which are an integral part of the boiler and setting, such as mechanical 
stokers, Dutch ovens, down-draft furnaces, and fire-tile combustion 
chambers incorporated with the regular setting. 

94. Hand-fired Furnaces. — Hand-fired furnaces, as a class, are most 
obstinate smoke producers. Although they can be operated efficiently 
without the production of objectionable smoke the result depends 
more upon the fireman than upon the design of the furnace. The chief 
difficulty with hand-fired furnaces lies in the intermittent nature of 
the firing. When a fresh charge of coal is fed into the furnace an enor- 
mous volume of volatile matter is evolved. For complete combustion a 



SMOKE PREVENTION, FURNACES, STOKERS 155 

corresponding amount of air must be supplied and intimately mixed 
with the volatile gases before contact is made with the comparatively 
cool heating surface. In the average hand-fired furnace the combustion 
chamber is so small and the heating surface is so close to the grate that 
the partly burned gases strike the heating surface before oxidation is 
complete and combustion is hindered or even completely arrested. 
The majority of so-called smoke preventers are merely devices for 
mechanically mixing the air and volatile gases. These include fire- 
brick piers, baffles, arches, and steam jets. There is no question as to 
the value of these mixing devices if properly installed, but the personal 
element is too variable a factor for consistent results and the ultimate 
solution lies in mechanical stoking. The most economical and smoke- 
less hand-fired plants are those that approach the continuous feed of 
the mechanical stoker. The following rules formulated by the Coal 
Stoking and Anti-Smoke Committee of the Illinois Coal Operator's 
Association for firemen using Illinois and Indiana coal in hand-fired 
furnaces give some idea of the principles involved in effecting smokeless 
combustion : 

1. Break all lumps and do not throw any in furnace any larger than 
one's fist. The reason for this is, that large lumps do not ignite promptly 
and their presence also causes holes to form in the fire, which allow the 
passage of too much air. 

2. Keep the ash pits bright at all times. If they become dark it is 
evident that the fire is getting dirty and needs cleaning, which, if not 
done, will cause imperfect combustion and smoke. If the furnace is 
equipped with a shaking grate, it should be operated often enough to 
prevent any accumulation of ashes in the fire. Do not allow ashes to 
collect in the ash pits, as they not only shut off the air supply, but may 
cause the grate to be burned. 

3. In firing do not land the coal all in one heap, but spread it over 
as wide a space as possible as it leaves the shovel. A little practice 
will enable one to catch the proper motion to give the shovel to make 
the coal spread properly. 

4. Place the fresh coal from the bridge wall forward to the dead plate 
and do not add more than 3 or 4 shovels at a charge. If this amount 
makes smoke it should be reduced till smoke ceases, which means, of 
course, that firing will be at more frequent intervals than formerly to 
keep up steam. This rule applies in cases where the boiler is worked 
at a large capacity. In such instances, however, where a small capacity 
only is required, firing by the coking method is the best, wherein the 
fresh coal is placed at the front of the fire and pushed back and leveled 
when it has become coked. 



156 



STEAM POWER PLANT ENGINEERING 



5. Fire one side of the furnace at a time so that the other side con- 
taining a bright fire will ignite the volatile gases from the fresh charge. 

6. Do not allow the fire to burn down dull before charging. If this 
is done, it will not only result in a smoky chimney, but an irregular steam 
pressure. 

7. Do not allow holes to form in the fire. Should one form, fill it by 
leveling and not by a scoop full of coal. Keep the fire even and level 
at all times. As far as possible level the fire after the coal has become 
coked. 

8. Carry as thick a fire as the draft will allow, but in deciding on the 
proper thickness, judgment must be exercised. If the draft is poor, a 
thin fire will be in order, but if strong, a thicker fire should be carried. 

9. Regulate the draft by the bottom or ash-pit doors and not by the 
stack dampers, because when the stack damper is used it tends to pro- 
duce a smoky chimney, as it reduces the draft, while the closing of the 
ash-pit door diminishes the capacity to burn coal. If strict attention 
is given to firing, and accounting to demand for steam, there will be no 
occasion to have recourse to dampers, except when there is a sudden 
interruption in the amount of steam being used. 

10. A good general rule is to fire little and often, according to steam 
demands, rather than heavy and seldom. The former means economy 
in fuel and a clean chimney, while the latter signifies extravagance in 
fuel and a smoky chimney. 



V//////////////////// //4. 




Fig. 83. Plain Dutch Oven. 



95. Dutch Ovens. — An independent furnace or Dutch oven in front 
of the boiler as illustrated in Fig. 83 provides one of the simplest 
methods of securing a large combustion chamber for the mingling of 



SMOKE PREVENTION, FURNACES, STOKERS 157 

the air and combustible gases before delivering them to the boiler proper. 
Such a furnace produces very high temperatures when operating under 
best conditions, and hence must be lined with fire brick of excellent 
quality. Although better than the ordinary setting the plain Dutch 
oven is too limited in length and capacity to prevent smoke from form- 
ing, except at very light loads. The velocity of the gases is usually 
too high to permit either a thorough mixture or complete oxidation 
before striking the boiler tubes. Steam jets placed at the sides of the 
setting and blowing across the fire are sometimes effective in mixing 
the air and combustible gases, but the best results are obtained by 
modifying the construction of the furnace to the extent of introducing 
deflecting arches which vary the direction of flow and by increasing the 
length of the path of the heated gases. The greater the length of the 
path and the greater the number of baffles the more thoroughly will the 
air and gases be mingled, but the intensity of draft will of course be de- 
creased in proportion. A compromise must therefore be made between 
required draft and length of path. The larger the extent of fire -tile 
surface the greater will be the regenerative effect, which is of particular 
importance in hand firing when the evolution of volatile gas is inter- 
mittent, but the first investment and cost of repairs and renewals are 
greater. There are little reliable data available pertaining to the rela- 
tion between capacitj^ of furnace and length of path of the heated gases 
for maximum efficiency. A modified Dutch oven is illustrated in 
Fig. 84. The extension front is not necessary with all types of boilers, 
as will be seen from Figs. 86 and 87, in which a tile roof and baffles 
suitably arranged within the setting proper simulate the Dutch-oven 
effect. 

96. Twin-Are Furnace. — This arrangement, illustrated in Fig. 84 in 
connection with a hand-fired return tubular boiler, is a double furnace 
formed by longitudinal arches extending between bridge wall and fire 
door. 

The furnaces are fed and manipulated alternately, the object being 
to have one furnace in a highly incandescent state, while green fuel is 
fed into the other. Air is admitted both below and above the grate, 
and the volatile gases are supplied with the necessary oxygen for com- 
bustion before they come into contact with the comparatively cool boiler 
surface. 

The gases from both furnaces first pass into a chamber formed by a 
single arch sprung across the entire inner setting from the side wall, a 
short retarding arch being placed between this intermediate chamber 
and the rear of the setting. A special tile of high-grade refractory clay 
is used, the thickness varying from 4 to 6 inches, depending upon the 



158 



STEAM POWER PLANT ENGINEERING 




SMOKE PREVENTION, FURNACES, STOKERS 



159 



size of furnace and the length of span. The furnace can readily be 
substituted for the ordinary types in common use under any standard 
tubular or water-tube boiler and may be installed either under the 
boiler, as indicated in the illustration, or in an extension Dutch oven. 
This is an excellent furnace, and when properly manipulated gives 
smokeless and efficient combustion. 

97. Chicago Settings for Hand-fired Return Tubular Boilers. — 
Figs. 86 and 87 show the general details of settings for return tubular 
boilers as recommended by the Chicago Department of Smoke Inspec- 
tion. The setting illustrated in Fig. 85, and known as No. 7, is ordi- 
narily installed on heating jobs where the rate of combustion is 12 pounds 
of coal per square foot of grate surface per hour or less, and those shown 



' i]f \ Jz^^^^^ ////^/////y///^ ^/ // y/y // / / '/// / /////////A 




m 



-30 ***«£ 



lllllllllllllllllllllllllirilllllllllllllTTTIIIIIIIIII 




Fig. 85. Smokeless Setting for Hand-fired Return Tubular Boiler. (Chicago Specifications.) 



in Figs. 86 and 87 where the rate of combustion is greater or where the 
plant has a regular power load. The dimensions in Fig. 85 refer to a 
specific set of conditions and are not general. All three settings require 
careful manipulation for smokeless combustion as is the case with hand- 
fired furnaces in general. It has been the experience of the Depart- 
ment that most violations of the smoke ordinance are due primarily to 
insufficient draft, the required rate of combustion being too high for 
the available air supply. The requirements outlined in paragraph 93 
apply equally well to these settings. The following specifications refer 
to Figs. 85, 86, and 87, the items in the specifications corresponding to 
the letters in the illustrations. 

A. Doors should be of a type allowing the admission of excess air 
over the fire when so desired. If panels are cut in the fire 
doors for this purpose, the aggregate area of the openings 
should be not less than 4 square inches to each square foot of 
grate surface. 



160 



STEAM POWER PLANT ENGINEERING 



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SMOKE PREVENTION, FURNACES, STOKERS 



161 




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162 STEAM POWER PLANT ENGINEERING 

B. Arches should be made of wedge brick or "bull heads" and not 

laid in two courses of 4j-inch brick. 

C. The bridge wall should be made of first-grade fire brick above 

the grate line and with fire-brick facing not less than 9 inches 
in thickness on the combustion-chamber side. The top row 
should be a row-lock course. Provision should be made in 
the building of the bridge wall for lateral expansion. 

D. The combustion chamber floor should be paved with fire brick 

laid on edge. 

E. Fire-brick lining below the arch skewbacks should be not less 

than 9 inches in thickness. Fire-brick lining above the arch 
system and behind the deflection arch may be 4| inches in first- 
grade fire brick, with headers every fifth row. 

F. Fire brick over firing-door liners should be arched. This rule 

also applies to brick above the clean-out door openings. 

G. Facilities for taking up arch thrust should be provided in every 

case by suitable metal reinforcements extending horizontally 
throughout the length of the arches. No air space should 
intervene between the metal reinforcements and the skew- 
backs. 

H. Herringbone or Tupper grates or other similar types should not 
be selected where bituminous coal forms the major portion of 
the fuel. 

I. The back arch is preferably sprung from side to side rather than 
from back wall to rear boiler tube sheet. No metal should be 
exposed to direct heat of gases. 

J. Chimney heights of less than 90 feet above the grate line should 
not be permitted, and this height allowed only when the 
chimney is direct connected to the boiler uptake. In case of 
a breeching and detached chimney, add to the height of chim- 
ney computed by standard methods (never less than 75 feet) 
10 feet for every turn of the breeching and one foot for each 
foot in length of the breeching. 

K. For boilers of all sizes, special provision for the examination of 
girth seams must be made. 

L. In the event of arch failures, the boiler should be immediately 
taken out of service. This is to avoid failure of the boiler 
shell due to heat being applied upon a portion of the heating 
surface over which a mud deposit has formed. 



SMOKE PREVENTION, FURNACES, STOKERS 



163 



98. Wooley Smokeless Furnace. — Fig. 88 shows a longitudinal sec- 
tion and a sectional plan of a Wooley smokeless furnace applied to a 
B. & W. boiler. The main features of the furnace are a dividing wall 
in the fire box and a deflecting wall in the combustion chamber. The 
dividing wall permits of the alternate method of firing, whereby one 
side of the furnace is always in an incandescent state while the other 



Uptake 




LONGITUDINALS.ECTION 




Fig. 



SECTIONAL PLAN 
Wooley Smokeless Furnace. 



side is being supplied with green fuel. If a mechanical stoker is used 
the wall in the fire box is omitted. The products of combustion are 
intended to be thoroughly mingled with the requisite amount of air by 
the deflecting walls before entering the regenerative or secondary com- 
bustion chamber. This type of furnace does not meet Chicago require- 
ments and is not much in evidence in the middle West. 

99. Kent's Wing-wall Furnace. — Fig. 89 shows the application of 
Kent's wing-wall furnace to a water-tube boiler. The Dutch oven 
in front of the regular setting contains the grates. Wing walls E are 
placed as shown two or three feet to the rear of the bridge wall D, and 
fire-brick piers H behind the wing walls. 



164 



STEAM POWER PLANT ENGINEERING 



In operation, fresh coal is spread alternately over each half of the 
grate. The dense smoky gases which rise from the green portion of 
the fire mingle in the narrow passage with the highly heated air which 
comes through the other side of the grate greatly in excess of that 




Fig. 89. Kent's Wing-wall Furnace. 

required to consume the partially burned coal there. The piers H act 
as regenerative surfaces, absorbing heat from the fire when it is hottest 
and giving it out when it is coolest, that is, just after firing. Wing-wall 
furnaces are little used in the middle West. 

100. Burke's Smokeless Furnace. — Figs. 90 and 91 show sections 
through a Burke smokeless furnace as installed in a number of tall 
office buildings in Chicago. It amounts virtually to a Dutch oven 
equipped with shaking grates, and embodies an extension self-feeding 
coking oven of cast-iron section lined with fire brick and protected from 
overheating by air circulation through the sections. Natural draft 
is used, the fire doors being closed; but air is admitted above as well 
as below the fire. As this stoker is manipulated by hand, more or 
less attention is required of the operator in keeping the fire clean. 
Furnaces of this type at the power plant of the Majestic Theatre build- 
ing, Chicago, 111., are giving excellent results. 

101. Down-draft Furnaces. — Fig. 92 shows the application of a 
Hawley down-draft furnace to a Heine water-tube boiler. In this 
furnace there are two separate grates, one above the other, the upper 



SMOKE PREVENTION, FURNACES, STOKERS 



165 



one being formed of parallel water tubes connected with the water 
space of the boiler through the steel headers or drums, A and D, in 
such a manner as to insure a positive circulation. Fuel is supplied 




Fig. 90. Burke's Smokeless Furnace — Front Section. 



BOILER 




*///////////////„ ,„„„„ „,„„/»////%. 

Fig. 91. Burke's Smokeless Furnace — Side Section. 



to the upper grate, the lower one, formed of common bars, being fed 
by the half-consumed fuel falling from the upper grate. Air for com- 
bustion enters the upper fire door, which is kept open, and passes first 
through the bed of green fuel on the upper grate and then over the 
incandescent fuel on the lower grate. A strong draft is required, due 



166 



STEAM POWER PLANT ENGINEERING 




SMOKE PREVENTION, FURNACES, STOKERS 167 

to the relatively small upper grate area and the correspondingly high 
rate of combustion. Lump coal gives better results than the smaller 
sizes, as the latter are apt to fall through the upper grate before being 
even partially consumed and when such is the case efficient results cannot 
be obtained. If carefully manipulated this furnace with fire-tiled tubec 
as illustrated in Fig. 92 gives high boiler efficiency and smokeless com- 
bustion, but its overload capacity is limited. Without the fire tiling 
smokeless combustion is possible only at light loads. 

102. Steam Jets. — The main purpose of a steam jet in connection 
with "smokeless furnaces" is to mix the air and gases and insure inti- 
mate mixture of the products of combustion. This action is purely 
mechanical, the steam in itself not being a supporter of combustion. 
The claims sometimes made that steam increases the calorific value 
of fuel are erroneous. There are conditions with certain grades of 
coals and refuse under which a moderate amount of steam injected into 
the furnace promotes complete combustion and increases the efficiency 
of the boiler. Such results, however, are due to increase in available 
heat and not to increase in actual calorific value. For example, hydro- 
gen and CO formed by the reaction between the steam and incandescent 
carbon unite with the oxygen of the air passing through the grate and 
generate intense heat. This heat dissociates a part of the steam into 
hydrogen and oxygen. The hydrogen immediately recombines with 
oxygen of the air, while the oxygen in its nascent state effects complete 
combustion of the hydrocarbons which under ordinary conditions escape 
in the form of smoke. Although it takes as much heat to dissociate 
steam into its elements as is given off when the hydrogen burns back 
again to water vapor the gain in available heat effected by the steam 
lies in the combustion of the hydrocarbons which would otherwise be 
discharged up the stack. The heat necessary to superheat the steam 
to stack temperature must be charged against the coal pile but the loss 
may be more than offset by this increase in available heat. It takes 
the same amount of oxygen to burn the hydrogen as is liberated by 
dissociation so there is no extra oxygen available for combustion, but the 
oxygen thus liberated is in a nascent state and combines much more 
readily with the hydrocarbons than does atmospheric oxygen. 

There is no question as to the value of properly installed steam jets 
in maintaining smokeless combustion in internally fired furnaces, hand- 
fired return-tubular boilers and improperly designed furnaces, but 
taking all things into consideration better results may be had with 
properly designed furnaces equipped with mechanical stokers. A 
smokeless stack with hand firing is not a true indication of efficient 
operation, since the air dilution may be excessive and the heat demands 



168 



STEAM POWER PLANT ENGINEERING 



of the steam jets may be very great. Since air requirements are greatest 
at the moment of firing fresh coal, and the demand diminishes as dis- 
tillation of the volatile matter progresses, steam jets need close regu- 
lation for best economy. If permitted to run continuously, as is often 
the case, they may use considerably more of the energy of the coal 
than they save by effecting smokeless combustion. In many of the 
patented " smoke consumers" the jets are automatic and operate inde- 
pendently of the fireman. 

Fig. 93 illustrates the application of a steam jet to a hollow bridge 
wall. The top of the wall is fitted with a small cast-iron column M, 
partially imbedded in the brickwork. A series of 1-inch holes "00," 
drilled near the top of the casting, furnish exits for the steam and air. 
A steam jet in one end of the column induces air into the iron chamber 
and forces it across the fire in fine streams. Excessive air dilution is 





GRATE 



Fig. 93. Applications of Steam Jets to Hollow Bridge Wall. 



avoided by partially closing the ash-pit doors and by regulating the 
intensity of the jets. An installation of this type is especially effective 
in connection with coal having a tendency to fuse and seal the air pas- 
sages in the grate. Two Stirling boilers at the Armour Institute of 
Technology equipped with this device gave practically smokeless com- 
bustion at all normal loads, though at heavy overloads it was necessary 
to slightly open the fire doors. Without the use of the jets smoke 
could not be prevented except at light loads. 

103. Parson Smokeless Furnace. — The Parson forced-draft system 
for smokeless combustion, applied to a return tubular boiler as illus- 
trated in Fig. 94, comprises a specially designed grate G, depending 
upon a steam jet blower A for draft. Part of the steam is admitted 
below the grate and part over the fire through the hollow bridge wall H. 
The supply of air above the grate is regulated by means of damper F. 
The steam to blower A is automatically adjusted by regulator A T , 
which is actuated by the steam pressure. The steam to the jet is super- 
heated by passing the supply pipe through the setting as indicated. 



SMOKE PREVENTION, FURNACES, STOKERS 



169 



The bridge wall H is provided with an extension platform M for hold- 
ing the unburned fuel when cleaning the fire. When equipped with a 
properly designed deflecting arch the furnace meets with the require- 
ments of the Chicago Smoke Department. 



Regulating VoiV.e 




9MS//S///M/>/S,_ 



Fig. 94. Parson Smokeless Furnace. 



104. Heinrich Smokeless Furnace. — Fig. 95 shows the application 
of the Heinrich system of forced draft to a return tubular boiler. 
Hot air is taken from the 
boiler room above the 
boilers by a steam jet 
blower at A and forced 
into the superheating 
chamber below the com- 
bustion chamber. From 
this chamber part of the 
air is drawn by the auxil- 
iary blowers C and forced 
through tuyeres above 
the grate, the rest passing 
through an opening be- 
neath the bridge wall 
into the ash pit and up 
through the bed of fuel. 
Steam for the blower A 
and the auxiliaries C is 
supplied through an auto- 
matic regulator R, which 




SECTIONAL PLAN 
Fig. 95. Heinrich Smokeless Furnace. 

opens when the steam pressure falls below the required value. 



170 



STEAM POWER PLANT ENGINEERING 



105. Luckenbach Smokeless Furnace. — Fig. 96 shows the applica- 
tion of the Luckenbach system to a standard return tubular boiler and 
setting. It consists primarily of a set of needle nozzles placed in the 

, ^ rrn side walls above the grate 

m 



and supplied with highly 
superheated steam from 
a specially constructed 
superheater. The steam 
jets are automatically 
controlled by the fire door 
and in such a manner 
that opening the door 
opens the jets and closing 
the door cuts off the 
steam supply after a 
predetermined period. 
The superheater consists 
Doo e r essentially of a wrought- 
iron coil cast within a 
solid block of iron and 
is placed on top of the 
bridge wall as indicated 
in the illustration. This 
superheater is an impor- 




SECTiabLAL PLAN 
Fig. 96. Luckenbach Smokeless Furnace. 



tant feature of the system. It is absolutely foolproof and does not 
burn out even if the steam supply is entirely cut off. (For further 
information see Eng. News, Dec. 29, 1910.) 

106. Mechanical Stokers. — Uniform evolution of the volatile gases of 
the fuel is the essential requisite for smokeless combustion, and it is 
for this reason that mechanical stokers as a class are more effective in 
preventing smoke than any apparatus accompanied by intermittent 
firing. Stokers which feed irregularly have the effect of hand-fired 
furnaces, and it is necessary not only to employ some powerful auxiliary 
mixing device but also to furnish at times an extra supply of air to take 
care of the enormous volume of volatile gas evolved after a fresh charge 
of fuel is added. 

Carefully adjusted automatic stokers owe their high efficiency to: 
(1) uniformity of feed; (2) proper proportion of air and combustible; 
(3) absence of exce'ssive air dilution, as when the fire doors are opened in 
connection with hand firing; and\4) self -cleaning grates. 

Daily records are essential with any type of stoker or hand firing 
if efficient results are expected, as only by frequent observation is it 



SMOKE PREVENTION, FURNACES, STOKERS 171 

possible to determine the proper adjustment of air supply, depth of fire, 
rate of feed, and the like. Control of air supply is almost as important 
as the upkeep and effective operation. In the best firing practice the 
right amount of air, depth of fire, and rate of feed must be worked out 
by the engineer. 

Stokers are often condemned by owners as inefficient and inferior to 
hand stoking because, no particular attention has been paid to them 
beyond filling the hopper with coal. They should be operated in strict 
accordance with the principles of design. 

In plants of 2000 horse power or over, the installation of mechanical 
stokers and coal conveyors effects a considerable saving of labor and 
can usually be relied upon to solve the smoke problem if reasonable 
attention is given to their operation. In smaller plants interest on 
the investment and other considerations may make hand firing more 
economical, although many plants of capacities as small as 200 horse 
power are giving satisfaction, particularly in places where a poor grade 
of fuel is used and smoke ordinances are rigidly enforced. A stoker 
of the self-cleaning, slow-running type requires much less attention 
than the hand-fired furnace. With hand firing one fireman can effi- 
ciently attend to the water, coal, and ashes of about 200 horse power, 
or handle coal for, say, 500 horse power, whereas with good automatic 
stokers he can readily take care of 2000 horse power or of 4000 horse 
power with chain grates equipped with overhead bunkers and down 
spouts. 

The best stokers are those which are least complicated and simplest 
in operation. A cheap stoker is a poor investment, since the cost of 
repairs and shutdowns will usually amount to more than the saving 
in price. 

The following outline gives a classification of a few of the best-known 
American mechanical stokers: 

Front Feed. 

Chain Grates: Step Grates: 
Babcock & Wilcox, Roney, 

Green, Wilkinson, 

McKenzie, Acme, 

Playford. McClave. 

Side Feed. Under Feed. 
Step Grates: Jones, 

Murphy, American, 

Detroit, Taylor, 

Model. Guckett. 



172 STEAM POWER PLANT ENGINEERING 

Down Draft. Sprinkler. 

Hawley. Swift, 



Powdered Fuel. g r - e 



Vulcan, 

See paragraphs 34-46. 

107. Chain Grates. — The chain grate is one of the most popular 
forms of automatic stokers for burning small sizes of free-burning 
bituminous coals. It embodies a moving endless chain of grate bars 
mounted on a frame with provision for the continuous and uniform 
feeding of coal into the furnace, the fuel and the grate moving together. 
As usually installed the surface of the grate is horizontal though in 
some designs it is given a slight incline toward the bridge wall. The 
operations of feeding the coal, carrying it through the progressive stages 
of combustion, removing the ashes and clinkers, and maintaining a 
clean grate and free air supply are automatic. The driving mechanism 
consists of a gear train actuated by ratchet and pawls, the arms carry- 
ing the latter being given a reciprocating motion by an eccentric mounted 
on a line shaft. The latter may be driven by any type of engine or 
motor and the speed of the grate regulated by varying the stroke of 
the arm carrying the pawls. Fuel is fed into a hopper placed at the 
front end of the furnace and the depth of the fuel regulated by a guillo- 
tine damper. The front part of the furnace is provided with a flat 
or slightly inclined ignition arch the function of which is obvious. The 
entire grate and driving mechanism are mounted on a permanent 
truck and may readily be removed from beneath the boiler. The 
thickness of the fire and the speed of the grate should be so regulated 
that when the fuel has reached the end of the grate it shall have been 
completely consumed and ashes only will be discharged into the pit. 
With chain-grate stokers there may be considerable leakage of air 
between the grate and bridge wall, through the coal in the hoppers, 
under the coal-gate and through the fire bed at the rear where it is 
mostly ashes unless care is used in regulating the depth of fire and ash 
bed and . provision is made for preventing this " short circuiting" of 
the air supply. 

Fig. 97 shows the general application of a B. & W. chain grate to a 
B. & W. boiler. The ignition arch is parallel to the grate and covers 
a considerable portion of the grate surface. The bridge wall is fitted 
with a water back as indicated, to prevent the grate bars from being 
burned. With uniform loads this style of ignition arch and setting 
insures practically smokeless combustion, but careful manipulation is 
necessary with rapidly fluctuating loads to prevent the formation of 
objectionable smoke. 



SMOKE PREVENTION, FURNACES, STOKERS 173 




Fig. 97. Babcock and Wilcox Boiler, Chain Grate, Ordinary Setting. 




Fig. 98. Babcock and Wilcox Boiler, Chain Grate, Fire-tile Roof. 



174 



STEAM POWER PLANT ENGINEERING 




SMOKE PREVENTION, FURNACES, STOKERS 



175 




176 



STEAM POWER PLANT ENGINEERING 



Fig. 98 shows the application of a fire-tile roof to a B. & W. boiler 
with chain grate. This arrangement, though effective in the preven- 
tion of smoke, is not much in evidence in modern plants because of the 
short life of the tiles. 

Fig. 99 shows the application of a B. & W. chain grate to the rear of 
a B. & W. boiler as installed at the Quarry Street Station of the Common- 
wealth Edison Company. Fig. 100 shows the application of a B. & W. 
stoker to the front end of the setting but with a much larger combustion 
chamber. Both of these settings are effecting smokeless combustion 
with Illinois screenings. 



•■^•••■r '->■••..-, ■■■ 




Fig. 101. Green Chain Grate, Standard Type. 

Fig. 101 shows the application of a standard Green chain grate for 
free-burning bituminous coals. The ignition arch is inclined to the 
grate and the distance between arch and grate is much greater than with 
either of the settings just described. The arch is inclined so that the 
heat from the incandescent fuel bed near the bridge wall is reflected 
down upon the green fuel, thus hastening distillation of the volatile 
gases. This arrangement permits of greater overloads with smokeless 
combustion. Swinging "air-back" E serves the double purpose of 
retaining the fuel and ash on the grate until all of the combustible is 
consumed and of preventing air leakage between the bridge wall and 
the end of the chain. The accumulated ash is dumped into the ash pit 
as occasion requires by rotating the air-back on its axis. A small steam 
jet or air blast prevents the air-back from being overheated. 



SMOKE PREVENTION, FURNACES, STOKERS 177 




178 



STEAM POWER PLANT ENGINEERING 



Coking bituminous coals are not adaptable to the ordinary style of 
chain grate on account of the swelling and fusing action of the coal 
under the ignition arch. Fig. 102 shows a Green grate modified to 
burn this grade of fuel. The fresh fuel is fed from the hopper onto 
dead plate B by means of pusher A and from the latter upon the in- 
clined coking plates, C, D. These plates gently agitate the fuel during 
the period of distillation and prevent it from fusing together so that by 
the time it reaches the grate proper, it no longer tends to cake. The 
ignition arch is inclined for the purpose previously stated. Grates of 
this type installed at the 59th Street station of the Interborough Rapid 
Transit Company are giving excellent results at heavy overloads. The 
results of these tests are shown in Fig. 57. 



MOPPEftEND 
>MKOAT-PL»fE \ 




Details of Construction of the Roney Mechanical Stoker 

Fig. 103. Details of Roney Stoker. 

108. Step Grates, Front Feed. — Fig. 103 shows the general arrange- 
ment of a Roney stoker and Fig. 105 that of a Wilkinson stoker, illus- 
trating the step-grate, front-feed principle. The Roney stoker consists 
of a hopper for receiving the coal, a set of rocking stepped grates inclined 
at a proper angle from the horizontal, and a dumping grate at the 
bottom of the incline for receiving and discharging the ash and clinkers. 
The dumping grate is divided into several sections for convenience in 
handling. The coal is fed onto the inclined grate from the hopper by 
a reciprocating " pusher" actuated by the " agitator. " The power is 
supplied through an eccentric operated by a small engine or motor. 
The normal feed is about 10 strokes per minute. The grate bars rock 



SMOKE PREVENTION, FURNACES, STOKERS 



179 



through an arc of 30 degrees, assuming alternately horizontal and 
inclined positions. The construction permits abundance of air to pass 
through the fuel, with little or no possibility of coal dropping through 
the grate. A coking arch of fire brick is sprung across the furnace as 
indicated. This stoker operates with natural draft and, with suitable 




Fig. 104. Double Stoker Installation at the 59th Street Station of the Interborough 

Rapid Transit Co., N. Y. 

arrangement of fire tiling, effects complete and efficient combustion. 
Without a fire-tile roof construction, smokeless combustion is effected 
with difficulty, particularly at heavy loads. 

In the Wilkinson stoker the inclined grate bars are hollow and are 
arranged side by side, every alternate bar being movable. When in 
operation there is a constant sawing action of the grate bars, causing 
the fuel to flow forward and downward. A small steam jet with about 
xVinch opening is introduced into the end of each hollow grate bar, 



180 



STEAM POWER PLANT ENGINEERING 



thus inducing the required amount of air for combustion, which passes 
through air openings approximately J-inch wide by 3 inches long. 
These stokers are driven by two small hydraulic motors, the water 
being furnished by a small pump and being used over and over again. 



THE MECHANISM OF THE WILKINSON STOKER 




Fig. 105. Details of Wilkinson Stoker. 

109. Step Grates, Side Feed. — Fig. 106 shows a front vertical- sec- 
tion and Fig. 107 a side vertical section through a Murphy automatic 
stoker and furnace. The apparatus is in effect a Dutch oven equipped 
with an automatic feeding and stoking device. Coal is introduced 
either mechanically or by hand into the magazine at each side of the 
furnace and above the grate and descends by gravity upon the coking 
plate. Reciprocating stoker boxes push the coal out upon the grate 
bars. Every alternate grate bar is movable and pivoted at its upper 
end. A rocker bar driven by a small motor or engine causes the lower 
ends to move up and down, this action producing the required stoking 
effect. A device for grinding up the clinker and ash is provided as 
shown at the bottom of the furnace. This is hollow and is connected 
by a 2-inch pipe with the smoke flue, so that the cold air passing through 
it prevents it being destroyed by the heat. Air is supplied to the green 
coal through flues passing under the coking plates, and the speed of 
the stoker boxes and grate bars can be regulated to conform to any 
rate of combustion. On account of the large fire-brick combustion 
chamber, this stoker with careful manipulation is capable of practically 
smokeless combustion. The power house of the Northwestern Elevated 



SMOKE PREVENTION, FURNACES, STOKERS 

v///m± 



181 




1UNSVERSE SECTION. 
Fig. 106. Murphy Furnace, Front Section. 




Fig. 107. Murphy Furnace, Side Section. 



182 



STEAM POWER PLANT ENGINEERING 



Railroad Company, Chicago, 111., is equipped with Murphy furnaces, 
which are operating smokelessly at an unusually high combustion rate, 
whereas a number of other installations using the same type of stoker 
and boiler and burning the same class of fuel are heavy smoke producers. 
Murphy furnaces are peculiarly adapted to variable loads, since at 
light loads the stoker may be operated with reduced grate area by 
allowing the bottom of the grate to partly fill with ashes. 

110. Jones Underfeed Stoker. — Fig. 108 shows the general principles 
of the Jones underfeed stoker. It consists of a steam-actuated ram 
with a fuel hopper outside of the furnace proper and a fuel magazine 
and auxiliary ram within. Air for combustion is admitted through 
openings in the tuyere blocks on either side of the retort. Coal is fed 
into hoppers and forced under the bed of fuel in the stoker retort, 
where it is subjected to a coking action. After liberation of the volatile 
gases the coke is pushed toward the top of the fire. The top of the 




Fig. 108. Jones Underfeed Stoker. 



fire, nearest the boiler, is always incandescent. Each charge of coal 
gives an upward and backward movement, which agitates the fire with- 
out opening fire doors. Air is admitted through the tuyere blocks at 
the point of distillation of the gases. 

Grate bars form no part of the Jones system, and it is therefore 
impossible for the fuel to fall through. There is no ash pit. The non- 
combustible matter is removed from the furnace by hand. The stand- 
ard size of the retort is about 6 feet in length, 28 inches in width, and 
18 inches in depth, and experience has shown that other sizes are not 
necessary since the spaces between retort and side wall of the various 
furnaces may be provided for by extending the width of the dead plates. 
One or more stokers are installed in each furnace, depending upon the 
capacity of the boiler and the width of the furnace. 

The steam pressure automatically controls air and fuel supply, pro- 
portioning them to each other and to varying loads in the correct 
degree. The result is that the stoker effects complete and smokeless 
combustion. The only variable element in the operation of this stoker, 
once it is correctly installed, is cleaning of fires, but if the fireman is 



SMOKE PREVENTION, FURNACES, STOKERS 



183 



careful to burn down the coals before breaking them up the production 
of smoke may be avoided. 

Jones underfeed stokers are adaptable to all grades and sizes of 
bituminous coal, and on account of forced draft are capable of burning 
very low grades of coal. 

111. American Underfeed Stoker. — Fig. 109 shows an application of 
an American underfeed stoker to a return tubular boiler. This differs 
from the Jones stoker in the method of feeding the fuel to the retort 
and in the employment of "live" grates instead of dead plates on the 
sides of the retort. The coal is fed into the hopper and carried by 

I ' ■ ■ - ■ / , , , TV 




Fig. 109. American Underfeed Stoker. 

an endless screw conveyor into the magazine or retort. Forced draft 
is used and the air supply and speed of the conveyor are readily adjusted 
to suit the conditions of load. In all underfeed stokers complete com- 
bustion is effected within a very short distance from the retort, hence 
a much smaller combustion space is required than with other types of 
stokers. For this reason underfeed stokers may give good results when 
installed in internally-fired boilers and small hand-fired return tubular 
boilers. 

112. Taylor Underfeed Stoker. — Fig. 110 shows the general details 
of a Taylor underfeed stoker for burning bituminous coals. The 
device consists essentially of a series of alternate retorts and tuyere 
boxes inclined as indicated. Each retort is fitted with two rams — 
the upper for pushing the green fuel outward and upward and the lower 
one for forcing the fuel bed and refuse toward the dump plates at the 



184 



STEAM POWER PLANT ENGINEERING 




SMOKE PREVENTION, FURNACES, STOKERS 



185 



rear. Air is supplied by a volume blower and enters the furnace through 
openings in the tuyere boxes. The dump plates are hung on the rear 
of the wind box and are controlled from the front of the stoker. Ex- 
tension grates are inserted between the mouth of the retort and the 
dump plates, when the nature of the fuel makes this arrangement desir- 
able. This extension may be rocked if necessary. The stoker and 
blower are operated by the same engine, the air and coal supply being 
automatically controlled by the variation in steam pressure. Taylor 
stokers may be operated smokelessly and efficiently at very heavy 
overloads and are much in evidence in the eastern states. The steam 
required to operate the blower and stoker varies from 2.5 to 5 per cent 
of the steam generated, de- 
pending upon the size of the 
installation and the percent- 
age of rating developed. 

113. Sprinkling Stokers. — 
In this system of stoking the 
fuel in finely divided form 
is distributed by sprinkling 
uniformly over the entire area 
of the grate. With the proper 
adjustment of air supply and 
feed the volatile gases are 
distilled off continuously 
before the grate is covered 
by the new coal and without 
materially lowering the tem- 
perature of the incandescent 
fuel. Mechanically the oper- 
ation involves considerable 
difficulty. Sprinkling stokers 
do not conform to Chicago 
requirements. 

Fig. Ill gives the general 
details of the Swift stoker, 
illustrating a commercially 
successful stoker of this type. 




Fig. 111. 



Swift Sprinkling Stoker. 



The apparatus is self-contained, and is bolted to a frame casting in front 
of the setting, and takes the place of the fire door. It may be swung 
back from the fire-door opening in much the same manner as the ordinary 
fire door. Coal of nut size or smaller is fed into a small hopper, of about 
300 pounds' capacity/from which it gravitates on to a berm plate and 



186 STEAM POWER PLANT ENGINEERING 

pusher plate. By means of the latter the fuel is fed to rapidly revolving 
spreaders, which crush it into small particles and throw it onto the grate. 
The fine or powdered coal is burned in suspension and the heavier coal 
falls to the grate. The spreaders are heavy pieces of cast steel, revolv- 
ing about a common axis and shaped helically so as to throw the fuel in 
a direction at right angles to the face of the machine. There are several 
of these spreaders so arranged on the shaft that adjacent spreaders 
throw the fuel in different directions. This stoker is not self-cleansing, 
that is, the ashes must be removed by hand or by suitable shaking 
grates. 

114. Smoke Determinations. — Smoke measurements may be either 
quantitative or relative. 

The most satisfactory method, at this writing, of determining the 
quantity of smoke passing through a chimney is that adopted by the 
Chicago Association of Commerce. A continuous sample of chimney 
gas is drawn from the stack by means of a special Pitot tube and ex- 
hauster, and the solid particles are entrapped in a filter. The tube is 
so arranged that the rate of flow through the apparatus is the same as 
that in the chimney. Since the area of the tube opening bears a fixed 
ratio to that of the chimney, the weight of carbon, cinders, soot and 
the like caught in the tube filter is a measure of the total weight emitted 
from the stack. 

Quantitative measurements are of considerable value in estimating 
the amount of energy lost in the production of visible smoke, but are 
seldom attempted in regular practice. 

There are several methods of determining smoke, relatively. The 
most common is that devised by Ringelmann, and is commercially 
known as the Ringelmann Smoke Chart. The chart, as published by 
the U. S. Geological Survey and used by the Smoke Department of 
the City of Chicago and other municipalities, consists essentially of a 
cardboard folder 12 by 26 inches over all. Four charts are printed on 
this folder, each chart consisting of 294 squares, 14 squares wide by 21 
squares in length, the width of the lines and spacings varying as illus- 
trated in Fig. 112. At a distance of 50 feet from the observer the lines 
become invisible and the cards appear to be of different shades of gray, 
ranging from very light gray to almost black. The observer places the 
chart on a level with the eye (at the distance stated, and as nearly as 
possible in line with the chimney) and notes which card most nearly 
corresponds with the color of the smoke. Observations should be 
made at 15-second intervals and recorded as in Fig. 113. No smoke is 
recorded as No. 0, 100 per cent as No. 5, and the intermediate colors 
as indicated by the cards. 



SMOKE PREVENTION, FURNACES, STOKERS 



187 



Experienced observers often record in half-chart numbers. Although 
these observations depend upon the personal element it is the opinion 
of the Chicago Smoke Department that only a little experience is 
necessary to effect consistent results with different observers. 

Prior to 1910 a chimney was held to be a smoke nuisance by the 
Chicago smoke inspection authorities when it emitted smoke of No. 3 



No. 



Fig. 112. 



No. 2. No. 3. No. 4. 

Ringelmann Smoke Chart (Greatly Reduced) . 



density, according to the Ringelmann chart, for 7 minutes during one 
hour, as based on the original ordinance. With this standard the 
owners of a chimney which emitted but a very small total quantity of 
smoke might be liable to punishment, whereas, with a chimney which 
continuously emitted smoke of a density less than No. 3, the owners 
would be safe from legal prosecution, although the total quantity 
emitted might be many times as great. 



43 5 2 4 4333 4 4 3 2 2 3322 22 322 




4 3 3 3 3 3 2 







| 1 1 










































































































i* 


1 


III, 


| [ 




~M 1 






























| 



















i 
A.M.. 


I 8 
50 


4 

i 


r 

5 


4 


6 

8. 


00 




8 




3 




5 







1 


2 



G 






34 
15 




4 




i 


5 





5 



Tim e 
Fig. 113. Smoke Record Chart. 

In future, the total smoke emitted will be taken into consideration. 
Observations will be made on a given stack every 15 seconds throughout 
the entire day and the total "smoke units" will be recorded, from 
which the average smoke density for the entire period will be cal- 
culated. 

A "smoke unit" is the equivalent of No. 1 smoke (Ringelmann 
scale) emitted for one minute. No. 1 smoke has a density of 20 per 



188 



STEAM POWER PLANT ENGINEERING 



cent; No. 2, 40; No. 3, 60; No. 4, 80, and No. 5, 100 per cent. Thus, 
if a stack emits No. 3 smoke for 6 minutes, 18 smoke units are charged 
against it. If this smoke was emitted during one hour's observation, 
then 

3 X 6 X 20 . 

xrr = b per cent 

is the average density of smoke emitted during the period of observa- 
tion. 

If observations on a given stack show that the density averages 
more than 2 per cent, although the owner may not be legally liable, 
an appeal is made to his personal and civic pride by a representative 
of the smoke-inspection department. For example, if a certain hotel 
stack emits smoke of more than 2 per cent average density, the smoke 
department finds a plant record of similar design and equipment, 
preferably a hotel plant, which shows a record well below the 2 per 
cent mark. This plant is then pointed out to the owner or manager 
having the objectionable chimney and he is asked if he cannot do 
equally well when he has practically the same equipment, etc. 

It has been found that this method of procedure often produces 
quicker and better results than a threat to go to law. 

New Methods of Approaching the Smoke Problem, Osborn Monnett, Jour. Wes. 
Soc. Engrs., Nov. 4, 1912. 

DIVISIONS OF MESH: RINGELMANN'S SMOKE CHART. 



Numbers give 

Relative Smoke 

Density. 


Thickness of 
Lines, mm. 


Distance in the 

Clear between 

Lines, mm. 




1 

2 
3 
4 
5 


All white 

1 

2.3 

3.7 

5 5 

All black 


All white 
9.0 
7.7 
6.3 
4.5 

i 



The Hammler-Eddy smoke recorder, Fig. 114, is one of the most 
successful devices for automatically recording the density of the smoke 
independent of personal observations. This apparatus consists essen- 
tially of a small motor-driven vacuum pump, which draws a continuous 
sample of the products of combustion from the uptake, breeching 
or stack and discharges it against a paper-covered drum revolved 
by clockwork. The density of the smoke, the time at which visible 
smoke is being emitted and the duration of the smoke-production period 
are automatically recorded on the paper by the smoke itself. Before 



SMOKE PREVENTION, FURNACES, STOKERS 



189 



reaching the pumps the gases pass through a glass " emergency" con- 
denser and a large portion of the vapor content is removed. The 
pump discharges the partially dried gases against a surface of sulphuric 
acid (which removes the last trace of moisture) and forces the smoke 
in the form of a small jet of dry powder onto the surface of the record- 
ing paper. The sampling tube leading from the flue to the pump is 




Fig. 114. Hammler-Eddy Smoke Recorder — Motor-driven Type. 

connected with a steam line and is "blown out" each time a card is 
changed. The instrument is very compact and portable and may be 
placed anywhere with respect to the chimney. A number of these 
appliances in Chicago power plants are giving excellent satisfaction. 
In a more recent design the pump is replaced by a steam siphon. 

115. Cost of Stokers. — The following is the approximate cost of 
stokers suitable for a Babcock and Wilcox boiler of 350-horse-power 
rated capacity with 45 square feet of grate surface; height of chimney 
above grate, 175 feet; coal burned, Illinois screenings. The cost of 
installation included, exclusive of brickwork, is 

1. Chain grate and appurtenances SI, 500.00 

2. Jones underfeed stoker 1,400.00 

3. Hawley down-draft furnace . 1,350.00 

4. Burke smokeless furnace ■ 1,000.00 

5. Roney stoker . . . 1,300.00 

6. Murphy furnace and stoker 1,350.00 

7. Wilkinson stoker 1,200.00 



CHAPTER V. 

SUPERHEATED STEAM; SUPERHEATERS. 

116. General. — The steam engine fails to realize the efficiency of 
the ideal engine chiefly on account of cylinder condensation. The loss 
in heat due to this cause is seldom less than 10 per cent of the total 
supplied, and often as great as 40 per cent. 

If the steam is superheated before being admitted to the cylinder, 
condensation may be reduced or prevented entirely, as was recognized 
as early as sixty years ago, but the mechanical difficulties encountered 
prevented the practice until within the past few years. 

The principal advantages of superheated steam in connection with 
steam-engine work are: 

1. At high temperatures it behaves like a gas and is, therefore, in a 
far more stable condition than in the saturated form. Considerable 
heat may be abstracted without producing liquefaction, whereas the 
slightest absorption of heat from saturated steam results in condensa- 
tion. If superheat is high enough to supply not only the heat absorbed 
by the cylinder walls but also the heat equivalent of the work done 
during expansion, then the steam will be dry and saturated at release. 
This is the condition of maximum efficiency in a single cylinder. (Rip- 
per, " Steam Engine Theory," p. 155.) Greater superheat than this 
will result in a loss of energy unless the steam is exhausted into another 
cylinder. To obtain dry steam at release the steam at cut-off must be 
superheated from 100 to 300 degrees F. above saturation temperature, 
depending upon the initial condition of the steam and the number of 
expansions, a higher degree of superheat being required for earlier cut- 
off. A superheat of from 250 to 350 degrees F. at admission is necessary 
to insure dry steam at release in the average single-cylinder engine 
cutting off at one-fourth stroke, boiler pressure 100 pounds gauge. In 
most cases superheat is only carried so far as to reduce initial conden- 
sation, the steam becoming saturated at cut-off, thus permitting efficient 
lubrication. There will be a reduction of approximately 1 per cent in 
cylinder condensation for every 7.5 to 10 degrees of superheat. In 
compound and 'triple-expansion engines the steam is ordinarily super- 
heated between each stage as well as before admission to the high- 
pressure cylinder. 

190 



SUPERHEATED STEAM; SUPERHEATERS 191 

2. A moderate amount of superheat produces a large increase in 
volume, the pressure remaining constant, and diminishes the weight 
of steam per stroke for a given amount of work. For example, the 
volume of one pound of saturated steam at 165 pounds pressure (abso- 
lute) is 2.75 cubic feet, and its temperature is 366 degrees F. The total 
heat of one pound of this steam above the freezing point is 11*95 B.t.u. 
By adding 108 B.t.u. in the form of superheat its temperature will 
be increased to 565.8 degrees F. (superheated 200 degrees F.) and its 
volume to 3.68 cubic feet (specific heat taken as 0.54). Thus an increase 
of 9 per cent in the heat effects an increase of 34 per cent in the volume, 
which means a corresponding reduction in the weight of steam admitted 
to the engine per stroke. These figures are purely theoretical, as no 
allowances have been made for condensation of the saturated steam 
or for reduction in temperature of the superheated steam. 

3. Superheated steam has a much lower thermal conductivity than 
saturated steam, and, therefore, less heat is absorbed per unit of time by 
the cylinder walls. With superheat smaller steam pipes may be used 
or a greater amount of power transmitted in a pipe of given size. By 
using high pressure and high superheat and then by lowering the pres- 
sure with reducing valves at the end of the line it is possible to transmit 
steam 10,000 feet or more without serious heat losses. 

117. Economy of Superheat. — Many comparative tests of engines 
using saturated and superheated steam under varying conditions of 
pressure and temperature have been made during the past few years, 
showing in all cases a gain in favor of superheat due to the reduction in 
steam consumption. In the majority of superheated steam installa- 
tions the ultimate gain is a substantial one, but in some cases the extra 
investment and cost of maintenance neutralize the reduction in steam 
consumption, resulting in an actual loss when measured in dollars and 
cents per horse-power hour. 

As far as steam consumption per horse-power hour is concerned, 
superheating usually increases the economy from 5 to 15 per cent and 
in some instances as much as 40, the latter figure referring to the more 
wasteful types of engines. A fair estimate of the average reduction 
in steam consumption per horse-power hour with moderate superheat- 
ing, that is, from 100 to 125 degrees F., based on continuous operation 
of existing plants, is: 

Per Cent. 

1. Slow running, full stroke, or throttling engines, including direct 

acting pumps 40 

2. Simple engines, non-condensing, with medium piston speed, includ- 

ing compound direct acting pumps 20 

3. Compound condensing Corliss engines 10 

4. Triple-expansion engines 6 



192 STEAM POWER PLANT ENGINEERING 

European builders guarantee steam consumption with highly super- 
heated steam as follows: 

Pounds per 
i.h.p. hour. 

Single-cylinder condensing engines (uniflow) 8.5 

Single-cylinder non-condensing engines (uniflow) 12.0 

Compound condensing engines (locomobile) 8.0 

Compound non-condensing engines (locomobile) 10.5 

The best recorded steam consumption at this writing (June, 1912) is 
that of a locomobile compound using steam superheated to 806 degrees 
F. at an initial pressure of 220 pounds absolute. When exhausting 
against an absolute back pressure of 1.32 pounds the steam consumption 
was 6.95 pounds per i.h.p. hour. (Zeit. des Ver. deut. Ingr., Mar. 18, 
1911, p. 415.) 

In comparing the performances of engines using saturated and 
superheated steam it is advisable to base all results on the heat con- 
sumed per horse power rather than on the steam consumption, since 
the latter is apt to give a false idea of the relative economies. The 
real measure of economy is the cost of producing power, taking into 
consideration all charges, fixed and operating, and the next best is 
the coal consumption per i.h.p. hour, but as a means of comparing the 
engines only, the heat consumption per horse power per hour or per 
minute is very satisfactory. (See paragraph 180.) 

See paragraph 208 for the influence of superheat on the economy of 
reciprocating engines and paragraph 235 for the influence on steam 
turbines. 

118. Limit of Superheat. — In this country steam temperatures ex- 
ceeding 500 degrees F. are seldom employed, while in Europe few if 
any plants are installed without superheaters, and 600 degrees F. is a 
common temperature with a maximum of about 850 degrees F. 

Experience has shown that with engines of ordinary design, slide- 
valves and Corliss, the temperature at the throttle should not exceed 
500 degrees F. This corresponds to a superheat of 160 degrees F. with 
steam at 100 pounds gauge pressure, and 130 degrees F. at 150 pounds. 
This degree of superheat insures practically dry steam at cut-off in the 
better grade of engines. Just how far superheating can be carried with 
a given engine of ordinary construction can be determined by experi- 
ment only, but a temperature of 500 degrees F. is probably an outside 
figure and 450 degrees F. a good average. Higher temperatures are 
apt to interfere with lubrication and sometimes cause warping of the 
valves. With temperatures below 450 degrees F. no difficulties are 
ordinarily met with. Metallic packing has been found to give the 
best results for both piston rods and valve stems. For highly super- 



SUPERHEATED STEAM; SUPERHEATERS 193 

heated steam " labyrinth" packing is used in place of the ordinary 
flexible metallic packing. 

It is generally assumed that a greater quantity of oil is required for 
lubricating valves and cylinders in connection with superheated steam, 
but experience seems to show that such is not the case. (Proc. A.S.M.E., 
May 14, 1908.) Forced-feed lubricators are the most satisfactory for 
superheated steam engines, since they insure a positive and copious 
flow of oil directly to the valves or other parts requiring it. (Effect 
of Superheated Steam on Cylinder Oils. Mech. Engr., Lond., July 31, 
1908, p. 115.) 

With highly superheated steam involving temperatures of 600 
degrees F. or more the poppet-valve type of engine (Figs. 220, 221) is 
ordinarily employed, though balanced piston valves are not uncommon. 
The poppet valve is not distorted by heat and requires no lubrication. 
In Europe these engines have been brought to a high state of efficiency, 
but have not been generally adopted in this country. 

119. Properties of Superheated Steam. — The laws governing super- 
heated vapors, like those governing saturated vapors, are not rational 
and deducible from a few fundamental experiments, but are more or 
less empirical in character. However, although the numerical values 
of the various quantities are based on the results of experiments, they 
permit of accurate mathematical formulation. Thus the following 
equations, derived by Prof . Goodenough (" Principles of Thermodynam- 
ics," 1911) and based upon the experiments of Knoblauch, Jakob, and 
Linde, give results which agree substantially with standard super- 
heated steam tables. 

T = absolute temperature of the superheated steam, degrees F. 
p = absolute steam pressure, pounds per square inch. 
X = total heat, B.t.u. per pound. 
u = intrinsic energy, B.t.u. per pound. 
n = entropy. 
C p = true specific heat. 
v = specific volume, cubic feet per pound. 
X = T (0.367 + 0.00005 T) - p (1 + 0.0003 p) ^ 

- 0.0163? + 886.7, (39) 

in which log C = 13.72511. 

u = T (0.2566 + 0.00005 T) - ^f- (1 + 0.00024 p) + 886.7, (40) 

in which log C = 13.64593. 



194 STEAM POWER PLANT ENGINEERING 

n = 0.8451 log T + 0.0001 T - 0.2542 log p 

- p (1 + 0.003 p) jrs ~ 0.3964, (41) 

in which log C = 13.64593. 
C p = 0.367 + 0.0001 T + p (1 + 0.0003 p) ^ > (42) 

in which log C = 14.42408. 

T 47 7Q^ V 10 9 

t; = 0.5963- - (1+ 0.0006 p) ' T , - 0.088, (43) 

in which log C = 13.64593. 

Wm. J. Goudie (Engng., July 1, 1910) gives the following simple 
equation for determining the specific volume which gives results suf- 
ficiently accurate for many engineering purposes. 

v = Vi (1 + 0.0016*), (44) 

in which 

Vi = specific volume of saturated steam, 

t = degree of superheat, degrees F. 

The mean specific heat may be obtained by subtracting the total 
heat of the saturated steam from that of the superheated steam and 
dividing the difference by the degree of superheat. 

Practically all commercial engineering problems are most con- 
veniently solved by means of superheated steam tables and diagrams, 
and recourse to formulas is seldom necessary. 

The curves shown in Fig. 115 give the true specific heats and those 
in Fig. 116 the mean specific heats of superheated steam for all pres- 
sure and temperature ranges likely to occur in practice. These curves 
are taken from Goodenough's " Principles of Thermodynamics," 1911, 
and are probably more accurate than those found in Marks and Davis' 
Steam Tables, though the difference is small. 

The Mollier diagram for superheated and saturated steam is repro- 
duced and described in Appendix L. 

A Discussion on Certain Thermal Properties of Steam: Prof. G. A. Goodenough, 
Jour. A.S.M.E., April, 1912. 

The Battle of the Superheats: R. H. Smith, Engineer, Dec. 22, 1911. 

Duchesne's Experiments on Superheat: V. Dwelshauver's-Dery, Power, July 23> 
1912, p. 110. 

Complete Discussion of the Work of Various Investigators: Dr. H. N. Davis, 
Proc. Am. Academy of Arts and Sciences, Vol. 45, p. 267, 1910. 

120. Superheaters. — Superheaters are manufactured by practically 
all boiler builders, the characteristics of the boiler being embodied to 
a large extent in the design of the superheater. The superheater may 



SUPERHEATED STEAM; SUPERHEATERS 



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196 



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SUPERHEATED STEAM; SUPERHEATERS 



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SUPERHEATED STEAM; SUPERHEATERS 199 

be independently fired or placed in the boiler setting. In the latter 
arrangement the superheater may be located between the furnace and 
the heating surface, as in Fig. 49, at the end of the heating surface, as 
in Fig. 125, or at some intermediate point, as in Figs. 119 and 121. 
Since the absorption of heat depends chiefly upon the average tem- 
perature difference between the gases and the steam and the extent of 
superheating surface, the required degree of superheat may be ob- 
tained from a small extent of heating surface in the furnace, a large 
amount in the rear of the heating surface or a proportionate amount in 
intermediate locations. In a general sense the sum of the boiler heat- 
ing surface and superheating surface per boiler horse power is prac- 
tically the same for any degree of superheat. The cost of a super- 
heated steam boiler is approximately equal to that of a saturated steam 
boiler since the superheated plant has less steam to generate. The 
requirements of a successful superheater are: 

1. Security of operation, or minimum danger of overheating. 

2. Economical use of heat applied. 

3. Provision for free expansion. 

4. Disposition so that it may be cut out without interfering with 
the operation of the plant. 

5. Provision for keeping tubes free from soot and scale. 

Superheaters may be separately fired or indirectly fired. The advan- 
tages of the separately fired superheater are: 

1. The degree of superheat may be varied independently of the per- 
formance of the boiler. 

2. It may be placed at any desired point. 

3. Repairs are readily made without shutting down the boiler. 

Some of the disadvantages are: 

1. It requires separate attention. 

2. Saturated steam only can be furnished to the prime movers in 
case of a breakdown to the superheater. 

3. Extra piping is required. 

4. Extra space is required. 

The indirectly fired superheater arranged in the boiler setting has 
the advantage of: 

1. Lower first cost. 

2. Higher operating efficiency. 

3. Minimum attention. 

4. Minimum space requirements. 



200 



STEAM POWER PLANT ENGINEERING 



As ordinarily installed the indirectly fired superheater is subject to 
the fluctuating temperatures of the furnace so that forcing the boiler 
has a similar effect on the superheater. In some cases the superheater 
adjusts itself automatically to the load requirements maintaining a 
constant degree of superheat at all loads, but in most cases the degree 
of superheat increases with the load, see Fig. 131. 

Standard central station practice in this country favors the super- 
heater contained within the boiler setting. 

121. Babcock and Wilcox Superheater. — Figs. 117 and 118 show the 
application of superheating coils to a Babcock and Wilcox boiler illustrat- 
ing the usual location of the indirectly fired type. The coils are made 




Fig. 117. Babcock and Wilcox Superheater. 



of 2-inch No. 8 gauge seamless steel tubes expanded into forged steel 
headers, the upper one receiving the saturated steam from the boiler 
and the lower one the superheated steam after it has traversed the 
superheater tubes. A small pipe connects the lower manifold with 
the water space of the boiler by means of which the superheater may 
be cut out if desired, or flooded when starting up. Any steam formed 
in the superheater tubes is returned into the boiler drum through the 
collecting pipe, which, when the superheater is at work, conveys sat- 
urated steam into the upper manifold. When steam pressure has been 
attained the superheater is thrown into action by draining the water 
away from the manifolds and opening the superheater stop valve. 
The tubes are free at one end and the manifolds are not rigidly con- 



SUPERHEATED STEAM; SUPERHEATERS 



201 



nee ted with each other, thus avoiding expansion strains. With the 
proportion of superheating surface to boiler surface ordinarily adopted 
the steam is superheated from 100 to 150 degrees F. 




Fig. 118. Babcock and Wilcox Superheater. 



122. Stirling Superheater. — This superheater consists of two drums, 
Fig. 120, connected by seamless drawn tubes two inches in diameter. 
It may take the place of the middle bank of tubes in the Stirling boiler, 
as shown in Fig. 119, or be installed in tjie final pass of the gases in the 
back of the boiler. The drums around the .tubes are protected from 
intense heat by asbestos cement. A pipe connecting the front drum 
of the boiler with the lower drum of the superheater permits the coils 
to be flooded in starting up or when the superheater is not needed. 
In this case the superheater acts as additional boiler-heating surface. 
The upper drum is divided into three and the lower into two compart- 
ments. The tubes are arranged with alternately wide and narrow 
spacing, so that any tube may be removed without disturbing the rest. 
The flow of steam is indicated by arrows. 

123. Foster Superheater. — Fig. 121 shows the application of a 
Foster superheater to a Babcock and Wilcox boiler. This device con- 
sists of cast-iron headers joined by a bank of straight parallel seamless 



202 



STEAM POWER PLANT ENGINEERING 



STEAM PIPE 



SAFETY. VALVE 




Fig. 119. Stirling Superheater. 




Fig. 120. Arrangement of Tubes; Stirling Superheater. 



SUPERHEATED STEAM; SUPERHEATERS 



203 



drawn-steel tubes, each tube being encased in a series of annular flanges 
placed close to each other and forming an external cast-iron covering 
of large surface. The protection afforded by this external covering is 
ample to prevent damage from overheating during the process of steam 



/Outlet- 




Fig. 121. Foster Superheater in Babcock and Wilcox Boiler. 

raising, and flooding devices are unnecessary. The tubes are double, 
the inner tube serving to form a thin annular space through which the 
steam passes as indicated. Caps are provided at the end of each ele- 
ment for inspection and cleaning purposes. Foster superheaters are 
more costly than plain-tube superheaters, but are longer lived and offer 
a much larger heating surface in proportion to the space occupied. 
Fig. 124 shows a Foster superheater arranged for independent firing. 
The "Schwoerer" superheater, which is somewhat similar in external 
appearance to the Foster, differs from it considerably in detail, the 
heating surface being made up of suitable lengths of cast-iron pipe 
ribbed outside circumferentially and inside longitudinally. The ends 
of the pipes are flanged and connected by cast-iron U-bends. Ine 
intention is to provide ample heating surface internally and externally, 
with a compact apparatus. 



204 



STEAM POWER PLANT ENGINEERING 








SUPERHEATED STEAM; SUPERHEATERS 205 

124. Heine Superheater. — Fig. 122 shows the application of a 
Heine superheater to a Heine boiler, illustrating the installation of a 
superheater within the boiler setting but entirely separated from the 
main gas passages. The superheater consists essentially of a number 
of lj-inch seamless steel tubes, bent to U-shape and expanded into a 
header box of the same type of construction as the standard Heine 
boiler water leg. The interior of this box is divided into three compart- 
ments by light sheet-iron diaphragms, so as to deflect the current of 
steam through the tubes. The superheater chamber is located above 
the steam drum as indicated. The gases of combustion are led to the 
superheater chamber through a small flue built in the side walls of 
the setting. A damper placed at the outlet of the flue controls the 
flow of gases and regulates the degree of superheat. No provision 
is necessary for flooding the superheating coils since the gases may 
be entirely diverted from the heating surface. Soot accumulations 
are readily removed by introducing a soot blower through the hollow 
stay bolts. 

125. Independently Fired Superheaters. — The Schmidt superheater, 
Fig. 123, consists of two nests of coils, A and D, of equal size and dimen- 
sions, connected to cast-iron headers and /. Saturated steam enters 
the first nest of coils through C and passes into header 0. From the 
steam, which is now dried, and partly superheated, flows through 
the cast-iron pipe E to header i", and thence through the second nest 
of coils into header adjoining 0, and through pipe R to the engine. 
In chamber D the steam and gases flow on the counter-current and in 
chamber A on the concurrent principle. This combination permits of 
a low flue temperature and high steam temperature without subjecting 
the tubes to an excess of heat as would be the case if the steam left the 
coils A at header i", where the furnace gases are the hottest. A steam 
temperature of 750 degrees F. and a flue temperature of 450 degrees F. 
are easily maintained with this apparatus. A mercury pyrometer T 
is fitted where the superheated steam enters the discharge pipe R. 
A thermometer cup L permits of checking the pyrometer by means of 
a nitrogen-filled thermometer. Each coil can be taken out separately 
and a new one put in without removing the others or dismantling the 
plant. Water produced by condensation while the superheater is 
inoperative collects in the bottom header N and escapes through a 
drain cock. If the steam supply should be suddenly shut off, the air 
door P is opened automatically by weight K. As soon as steam begins 
to flow it raises the weight through the opening of valve C and the 
door closes. The Schmidt superheater when arranged in the flue has 
practically the same construction as the independently fired. 



206 



STEAM POWER PLANT ENGINEERING 




SUPERHEATED STEAM; SUPERHEATERS 



207 




1 



I 
d 



208 



STEAM POWER PLANT ENGINEERING 



Fig. 125 shows a combination of Schmidt superheater, economizer, 
and feed-water heater which finds much favor with engineers on the 
continent. 




Fig. 125. Schmidt System of Combined Superheater, Feed-water Heater, and 

Economizer. 



126. Luckenbach Superheater. — Fig. 126 shows a section through a 
Luckenbach superheater illustrating an extremely simple and effective 
device for superheating small quantities of steam up to very high 
temperatures. It consists essentially of a single coil of extra heav}^ 
lj-inch steel pipe imbedded within the walls of a cylindrical casting. 
The coil is not welded to the casting but is free to expand and contract 
independently. The furnace illustrated in Fig. 126 is designed for 
hard coal or coke. The apparatus is compact and durable and no 
provision is necessary for flooding the coils. The steam supply may be 
cut off entirely with a furnace full of incandescent fuel without burning 
out the coils. The following results were obtained from capacity tests 
of a Luckenbach superheater 30 inches in diameter by 20 inches in 
height, as installed in the Mechanical Engineering Laboratory of the 
Armour Institute of Technology. 



SUPERHEATED STEAM; SUPERHEATERS 



209 



CAPACITY TEST OF A 30-INCH LUCKENBACH SUPERHEATER. 

Lineal feet of superheater coil 28 feet. 

Internal heating surface of furnace walls 7.5 square feet. 

Grate area 1.5 square feet. 



Steam Pressures, Lbs. 
per Sq. In. Gauge. 


Moisture 
in Steam 
Entering 
Superheater, 
Per Cent. 


Steam Temperat 
Degrees F. 


ires, 


Weight of 

Steam 

Flowing, 

Lbs. per 

Hr. 


Weight of 
Coke 


Entering 
Super- 
heater. 


Leaving 
Super- 
heater. 


Entering 
Super- 
heater. 


Leaving 
Super- 
heater. 


Degrees of 
Super- 
heat. 


Fired, 

Lbs. per 

Hr 


50 


40 


0.9 


298 


587 


300 


739.6 


17.7 


50 


45 


0.9 


298 


618 


325 


678.5 


17.1 


50 


46 


1.0 


298 


648 


354 


585.4 


17.1 


50 


46.5 


1.1 


298 


700 


406 


505.5 


17.6 


50 


47 


2.3 


298 


760 


465 


371.0 


17.6 


50 


47.5 


1.4 


298 


790 


495 


341.0 


17.7 


70 


67 


0.7 


316 


803 


490 


359.0 


17.7 


70 


66 


1.6 


316 


705 


392 


484.0 


*10.2 



Damper wide open throughout all tests, no attempt being made to obtain high furnace efficiency. 
* Economy test, damper throttled. 

127. Materials used in Construction of 
Superheaters. — Most superheaters are 
constructed either of wrought iron, mild 
steel, cast iron, or cast steel, the latter 
material having the advantage of not 
being damaged by any temperature to 
which it is likely to be subjected, which 
does away with the necessity of damper 
mechanisms and simplifies the installation. 
On the other hand, cast-metal superheaters 
are usually ribbed after the fashion of an 
■air-cooled gas engine, and are, therefore, 
very heavy and thick walled, necessitating 
a higher temperature for the same useful 
effect than in the case of the wrought- 
iron construction, but have the advantage 
of minimizing fluctuation of steam tem- 
perature which would otherwise be caused 
by a wide variation in temperature of 
furnace. One of the most successful cast- 
metal heaters is of European design and 
is constructed of a special alloy known as 
"Schwoerer" iron. Table 36 gives the yearly cost of repairs to piping 
and necessary brickwork for a number of installations equipped with 
cast-metal superheaters of the "Schwoerer" type. 




<*^<^ ^ i «■ " <■ 



ssW^a 



Fig. 126. 
Section through Luckenbach Inde- 
pendently Fired Superheater. 



210 



STEAM POWER PLANT ENGINEERING 



80,000 




Fig. 127. 



Temperature, Degrees Fahrenheit 

Effect of Temperature on Strength 
of Materials. 



Wrought iron and mild steel offer the advantage of lightness, ease of 
construction, and low first cost, but cannot be exposed to very high 
temperatures without injury, and consequently provision must be made 
for diverting the direction of the heated gases or for flooding the coils 
while the boiler is being warmed before steam is generated. 

The effect of temperature on 
superheater materials is shown in 
Fig. 127. It will be seen that the 
tensile strength drops off very 
rapidly for temperatures beyond 
650 degrees F. Because of this 
rapid decrease in tensile strength 
of materials with the increase in 
temperature, steam is seldom 
superheated to temperatures 
above 850 degrees F. 

For further information per- 
taining to the effect of tempera- 
ture on various metals, consult 
"The Effect of High Tempera- 
tures on the Physical Properties of Some Metals and Alloys"; The 
Valve World, Jan., 1913, published by the Crane Co., Chicago. 

Ordinary cast-iron valves and fittings have shown permanent in- 
crease in dimensions under high superheat and in numerous instances 
have failed altogether, but sufficient data are not available to prove 
conclusively the unreliability of cast iron if the iron mixture is properly 
compounded and the necessary provision is made for expansion and 
contraction. Authorities are of the opinion that the failure of cast- 
iron fittings is due more to fluctuations in temperature than to the 
actual high temperature itself and cite numerous cases where ordinary 
cast-iron fittings under uniform temperature conditions are giving satis- 
faction with highly superheated steam. Notwithstanding the claims 
that cast iron properly compounded is a perfectly reliable metal for 
fittings, engineers are inclined to use cast or forged steel, at least in 
this country. See "Effect of Superheated Steam on Cast Iron and 
Steel," Trans. A.S.M.E., Vol. 31, 1909, p. 989. 

128. Extent of Superheating Surface. — The required extent of super- 
heating surface for any proposed installation depends upon: (1) the 
degree of superheat to be maintained; (2) the velocity of the steam 
and gases through the superheater; (3) the character of the super- 
heater; (4) the weight of steam to be superheated; (5) the moisture 
in the wet steam; (6) the temperature of the gases entering and leav- 



SUPERHEATED STEAM; SUPERHEATERS 



211 



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212 STEAM POWER PLANT ENGINEERING 

ing the superheater; (7) the conductivity of the material, and (8) clean- 
liness of the tubes. 

Since the heat absorbed by the steam in the superheater is equal to 
that given up by the products of combustion, neglecting radiation, this 
relationship may be expressed 

SUd = Wcih-U), (45) 

in which 

S — square feet of superheating surface per boiler horse power, 

U = coefficient of heat transmission, B.t.u. per hour per degree 

difference in temperature, 
d = mean temperature difference between the steam and heated 

gases, degrees F., 
W = weight of gases passing through the superheater per boiler- 
horse-power hour, 
c = mean specific heat of the gases, 

ti = temperature of the gases entering superheater, degrees F., 
t 2 = temperature of the gases leaving superheater, degrees F. 
Transposing equation (45), 

S-- m . (46) 

The heat transfer from the products of combustion to the steam may 
also be expressed 

SUd = wc' (t a - t), (47) 

in which 

w = weight of steam passing through the superheater, pounds per 

boiler-horse-power hour, 
c' = mean specific heat of the superheated steam, 
t, = temperature of the superheated steam, degrees F., 
t = temperature of the saturated steam, degrees F., 
S, U, and d as in equation (45). 
For wrought-iron or mild steel tubes U varies as follows: 
U = 1 to 3 for superheaters located at the end of the heating surface, 
= 3 to 5 for superheaters located between the first and second 

pass of water tube boilers, 
= 8 to 12 for superheaters located immediately above the fur- 
nace in stationary boilers, in the smoke box of locomotive 
boilers, and in separately fired furnaces. 
General practice allows J to } square foot of superheating surface per 
boiler horse power for mild steel, superheater located in the furnace; 
from 2 to 2.5 square feet of surface at the end of the first pass, and 
from 3 to 4 square feet at the end of the heating surface for superheats 
of from 100 to 150 degrees F., boiler pressure 150 pounds absolute. 



SUPERHEATED STEAM; SUPERHEATERS 213 

The Foster Superheater Company allows 6 B.t.u. per lineal foot per 
degree difference in temperature for their "two-inch" element where 
the average temperature of the gases is about twice the mean temper- 
ature of the steam. 

For all engineering purposes d may be determined with sufficient 
accuracy from the relationship 

, £1 + £2 t a + t 
d = ~2 2~- 

Notations as in equations (45) and (47). 

An empirical formula for determining the extent of superheating 
surface in connection with indirect superheaters which appears to 
conform with practice is derived by substituting 

U = 3, d = t' - t ^- t , w = 30, c' = 0.5, 

in equation (47) [J. E. Bell, Trans. A.S.M.E., 29-267]. Thus: 
S X 3^' - ^±-^ = 30 X 0.5 X (t. - t), 



from which 



_ 10 «. - 1) . 

S -2t'-t.-t' (48) 



t' (the mean temperature of the product of combustion where the super- 
heater is located) may be approximated from equation 

(< >_ 1 .. M = 0.172fl + 0.294, (49) 

in which 

H = the per cent of boiler-heating surface between the point at 
which the temperature is t and the furnace, 
t as in (48). 

Equation (49) is based upon the assumption that the heat trans- 
ferred from the gases to the water is directly proportional to the differ- 
ence in temperature; that the furnace temperature is 2500 degrees F.; 
flue temperature 500 degrees F.; steam pressure 175 pounds per square 
inch gauge; one boiler horse power is equivalent to. 10 square feet of 
water-heating surface. 

Example: What extent of heating surface is necessary to superheat 
saturated steam at 175 pounds gauge pressure, 200 degrees F., if the 
superheater is placed in the boiler setting where the gases have already 
traversed 40 per cent of the water-heating surface? 



214 



STEAM POWER PLANT ENGINEERING 







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SUPERHEATED STEAM; SUPERHEATERS 



215 



Substitute H = 0.4 and t = 378 in equation (49), 

y_ 3 78r 6 = 0.172X0.4 + 0.294, 

from which t' = 950. 

Substitute t' = 950, t 8 = 578, and t = 378 in equation (48), 

10 (578 - 378) 
2 X 950 - 578- 378 
= 2.12 square feet. 

The curve in Fig. 128 was plotted from equation (49) and gives a 
ready means of determining t' and of observing the law governing heat 
absorption by the boiler between furnace and breeching. The abscissas 
represent the temperatures of the hot gases at different points in their 
path between furnace and breeching. The ordinates represent (1) the 
per cent of boiler-heating surface passed over by the hot gases, and 
(2) the per cent of the total heat generated which is absorbed by this 
heating surface. 

In the use of equation (49) the probability of error is greatest when 
considering a point near the furnace, since large quantities of heat are 
transmitted to the tubes by radia- 
tion from the fuel bed which are not 
taken account of. For most practical 
purposes the assumption is suffi- 
ciently accurate. 

Fig. 129 gives the probable tem- 
perature range of gases entering 
superheater after passing over a 
given per cent of boiler-heating sur- 
face and Fig. 130 shows the relation 
between superheating surfaces and 
boiler heating surface. (See Power, 
Nov. 7, 1911, p. 696.) 

It will be found that the boiler- 
heating surface per boiler horse 
power will be decreased in almost the same proportion that the super- 
heating surface is increased, so that the sum of the boiler-heating surface 
and superheating surface per boiler horse power will be very nearly 
the same for any degree of superheat. 

For the application of the curve in Fig. 128 to the design of direct 
and indirect superheaters for various degrees of superheat, see " Stirling, " 
published by the Stirling Boiler Company, pp. 92-96. 




Fig. 129. 



55 50 

Per Cent 
Boiler Heating Surface Used 
before Reaching Superheater 

Temperature Range of Gases 
in Superheater. 



216 STEAM POWER PLANT ENGINEERING 

129. Performance of Superheaters. — Published tests of both directly 
and indirectly fired superheaters cover such a wide range of conditions 
of installation and operation that general conclusions cannot be drawn, 
but it may be of interest to note briefly the performances in a few 
specific cases. 

The curves in Figs. 131, 132, and 133 are plotted from tests of a Bab- 
cock and Wilcox boiler, with 5000 square feet of water-heating surface, 
equipped with superheating coils of 1000 square feet area, as illustrated 
in Fig. 97. The furnace with ordinary short ignition arch was pro- 
vided with chain grate of 75 square feet area. 

Fig. 131 shows the relation between degrees of superheating and 
total horse power of boiler and superheater. 

Fig. 132 shows the relation between horse power produced in the 
boiler and the percentage of boiler horse power produced in the super- 
heater. 

Fig. 133 shows the relation between the degree of superheat obtained 
and the horse power developed in the superheater. 

Tables 37 to 39 are taken from the report of Otto Berner ("Zeit. d. 
Ver. Deut. Eng." and reprinted in Power, August, 1904). 

Table 37 compares the heat efficiency of a steam plant equipped with 
directly and with separately fired superheaters, the former showing a 
much higher efficiency. 

Table 38 compares different boilers with and without flue super- 
heaters, showing the effect upon the temperature of the flue gases. 
The gain in heat efficiency of the entire plant due to the use of the 
superheater is decisive in each case. 

Table 39 shows the gain in heat efficiency due to the use of super- 
heaters in a number of plants equipped with fire-tube boilers. 

Table 40 gives the results of tests on one of the return tubular boilers 
at the Spring Creek Pumping Station of the Brooklyn Waterworks 
(Feb. 9, 1904) with and without a superheater. The superheater, of 
the Foster type, was installed between the rear wall of the setting and 
the tube sheet. 

Although the results in Tables 37 to 40 represent practice of eight 
years ago, they agree substantially with current practice (1912). 



SUPERHEATED STEAM; SUPERHEATERS 



217 



& 



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400 











































































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10 



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20 25 30 
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Fig. 130. Relation between Superheat and 
Boiler Heating Surface. 



u £ 



£20 



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200 



400 



600 



800 



Horse Power Produced in Boiler 

Fig. 132. Ratio of Horse Power produced in 
the Superheater to that developed in the 
Boiler. 



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Degrees in Superheat— ~F 
Fig. 131. Relation of Degree of Superheat 
to Total Horse Power developed. 



120 
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w 50 100 150 200 

Degrees of Superheat~F 

133. Relation of Degree of Superheat to 
Horse Power of Superheater. 



218 



STEAM POWER PLANT ENGINEERING 



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SUPERHEATED STEAM; SUPERHEATERS 



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SUPERHEATED STEAM; SUPERHEATERS 



221 



TABLE 40. 

{Engineer, U. S., May 1, 1904.) 



Time of start 

Time of finish 

Hours run 

Average steam pressure 

Average water pressure, triple expansion 

head in feet 

Average water pressure, compound, head 

in feet 

Average vacuum of suction for triple and 

compound, inches of mercury 

Total head on triple, feet of water 

Total head on compound, feet of water . . . 

Total double strokes, triple 

Total double strokes, compound 

Gallons pumped from piston displacement, 

total, triple 

Gallons pumped from piston displacement, 

total, compound 

Gallons pumped from piston displacement, 

total, triple combined 

Gallons, total, pumped as measured by weir 

Per cent slip 

Foot pounds, weir 

Total coal consumed 

Per cent refuse 

Total refuse 

Total feed water 

Duty per 100 pounds coal 

Duty per 1,000 pounds steam 



With Superheater. 


Without Superheater 


12 noon, Feb. 8 
12 noon, Feb. 9 


11 a.m., Feb. 11 
11 a.m., Feb. 12 


24 
79.3 1b. 


24 
79.4 1b. 



0.99 



7.10 



22.90 


23.21 


29.05 


29.46 


33.04 


33.39 


30,557 
35,395 


34,114 
32,158 


2,854,023 


3,186,247 


2,930,706 


2,662,682 


5,784,720 
4,492,680 
22.3 


5,848,930 
4,549,480 
22.2 


1,163,815,819 

5,015 lb. 

23.7 


1,184,983,596 

6,410 lb. 

18.7 


1,188 

38,399 

23,206,696 

30,308,498 


1,203 

50,960 

18,486,483 

23,253,213 



1.05 



7.10 



Per cent increase of work per 100 pounds coal 25.5 

Per cent increase of work per 1,000 pounds steam 30 . 2 

Per cent saving in coal per foot pound work 20 . 2 

Per cent saving in feed water per foot pound work 23 . 2 

Average temperature steam leaving superheater 527. 4 deg. F. 

Average temperature steam entering superheater 320. 1 deg. F. 

Average degree superheat 207.3 deg. F. 



CHAPTER VI. 

COAL AND ASH-HANDLING APPARATUS. 

130. General. — The cost of coal and its delivery into the furnace 
are usually the largest items in the operating charges; hence large 
central stations are located, when practicable, adjacent to a railway 
line or water front, to minimize the cost of handling coal and ashes. 
Isolated stations in the business districts of large cities are usually 
unfavorably situated, so that the cost of handling coal and ashes is a 
large percentage of the total fuel cost. In large stations the amount 
of fuel and ash handled frequently warrants the expense of elaborate 
conveyor systems which would not be justified in smaller plants. In 
whatever way coal is supplied provision should be made for storing a 
quantity sufficient to operate the plant for some time in case the supply 
is interrupted, thereby guarding against an enforced shut-down. 

If adjacent to a railway line, a side track must be provided for switch- 
ing the cars. As bottom-dumping cars cannot be depended upon, pro- 
vision should be made for unloading by hand or by grab bucket. If 
coal is delivered by water, clam-shell drop buckets are ordinarily used 
for unloading the barges. If the power house is located at some dis- 
tance from the railroad or water the coal is generally hauled by teams in 
two- to five-ton loads. 

131. Coal Storage. — In small stations the storage bins or coal 
bunkers may usually be located within the building, but in larger plants 
the quantity of coal consumed daily is frequently such that an immense 
space would be required to furnish storage capacity for even a short 
period of time. For example, one of the large central stations in 
Chicago burns an average of 60 tons of Illinois screenings per hour 
throughout the year. Allowing 45 cubic feet to the ton this would 
necessitate a space of 45 X 60 X 24 = 69,600 cubic feet to store coal 
for one day's operation. A ten-days' run would require a coal pile 50 
feet wide, 30 feet high, and 464 feet long. It is a good plan, if the 
location and character of the plant permit, to carry four or five days' 
supply within the plant and provide a separate building for the coal 
reserve. Such provision is made in the power plant of the New York 
Edison Company, which has a storage capacity of 150,000 tons in ad- 
dition to that of the overhead bunkers. 

222 



COAL AND ASH-HANDLING APPARATUS 223 

Exposed coal piles are objectionable, because of freezing in winter, 
the crust sometimes freezing so hard as to necessitate the use of dyna- 
mite to break it; moreover, a slow depreciation in heat value takes 
place, especially with bituminous coal. This depreciation is more rapid 
in warm weather and in the tropics. Stored coal is oftentimes subject 
to spontaneous combustion, particularly when there is a large content 
of iron pyrites. Storage under water minimizes spontaneous combus- 
tion and depreciation in heat value. (Consult references below.) 

Coal bunkers or hoppers are ordinarily placed on the same level with 
the boiler-room floor or above the boiler setting. The former is the 
cheaper as far as first cost is concerned, but necessitates additional han- 
dling of the fuel before it can be fed to the stokers. In the overhead 
system the coal gravitates to the stoker through down spouts. Over- 
head bunkers are usually found where real estate is costly. They are 
generally constructed of steel plates lined with concrete or of reenforced 
concrete. The bottoms slope at an angle of 35 to 45 degrees and 
empty into the coal chutes or down spouts. Fig. 136 shows the general 
appearance of a single overhead bunker. In some bunkers the floors 
are made with very slight slopes, but it is not advisable to use a 
slope less than the angle of repose of the coal, as it may be necessary 
to shovel the coal over the spouts. Convenience in framing makes the 
45-degree slope the more desirable. Separate bunkers for each boiler 
are preferred to continuous bunkers, since fire in the coal is more 
readily prevented from spreading. In the new power house of Swift 
& Co., Chicago, 111., the bunkers are of circular cross section instead of 
rectangular, as is the usual practice. The capacity of the cylindrical 
hopper is considerably less than that of a rectangular hopper of the 
same width, but is much cheaper to construct. 

Ash bins are invariably lined with concrete or brickwork, since the 
corrosive action of the ashes would soon destroy the bare iron, and are 
usually located alongside the coal hopper, as in Fig. 136, so that they 
may be discharged by gravity. The angle of repose of most ashes 
is approximately 40 degrees, but the 45-degree angle is preferred on 
account of convenience in construction. Fig. 102 illustrates a " non- 
arching" type of ash hopper in which the sides are flared sufficiently 
to prevent the ash from packing when the bottom valves are opened. 

Coal Storage Under Water: Elec. Wld., Oct. 7, 1911, p. 885, Eng. News, Dec. 24, 
1908. 

Calorific Value of Weathered Coals: Bulletin No. 17, Univ. of 111., Aug. 26, 1907; 
Eng. News, Jan. 11, 1912, p. 64. 

Spontaneous Combustion of Coal: Jl. Ind. and Chem. Eng., Mar., 1911. 

Suspended Coal Bins: Power, Apr. 23, 1912, p. 602. 



224 STEAM POWER PLANT ENGINEERING 

132. Coal Conveyors. — Coal is carried to the stokers in a variety of 
ways, depending upon the location of the plant, the type of stokers, 
and the personal tastes of the builder. Of the various methods the 
following are the most common: 

1. Hand shoveling from coal pile to furnace. 

2. Wheelbarrow or hand car and shovel. 

3. Bucket conveyor. 

4. Belt conveyor. 

5. Hoist and hand cars. 

6. Hoist and automatic-cable cars. 

7. Hoist and trolley. 

8. Spiral or screw conveyor. 

9. Combinations of the above. 

For a series of papers on conveying machinery with data pertaining 
to the cost of operation see Trans. A. S. M. E., Vol. 30, 1908, and Indus- 
trial Engineering, Oct., 1911, to March, 1912 (Serial). 

133. Hand Shoveling. — Where possible, the coal is dumped direct 
from the cars or wagons into bins located in front of the boilers. In 
such instances one man may handle the coal and ashes and attend to 
the water level of 200 horse power of boilers equipped with common 
hand-fired furnaces. With stoking and dumping grates 300 horse 
power may be controlled by one man and from 800 to 1000 horse power 
with chain-grate stokers. This refers, of course, to average good coal 
not too high in ash nor productive of much clinker. Sometimes the 
coal cannot be stored in front of the boilers but must be hauled by 
wheelbarrow, cart, or rail car. For distances over 100 feet and quanti- 
ties over 20 tons per day the cost of handling the coal in this way may 
justify the installation of an automatic conveyor system. Hand-fired 
furnaces and manual handling of coal and ashes are usually associated 
with small plants of 500 horse power and under, but a number of large 
stations are operated in this way with apparent economy. A notable 
example is the new (1907) steam power plant of the Wood Worsted 
Mill, Lawrence, Mass., in which 40 return tubular boilers are fired by 
hand. A tipcart with a capacity of one ton brings the coal a distance 
of 100 to 200 feet to the firing floor, and firemen shovel it on to the 
grate. Four men are stationed at the coal pile. One man drives two 
carts (one of which is being filled while the other is gone with its load), 
sixteen firemen attend to the furnaces, and two men dispose of the 
ashes. Most large plants, however, are equipped with conveying 
machinery, not so much because of the possible reduction in cost of 
operation, taking into consideration all charges fixed and operating, as 



COAL AND ASH-HANDLING APPARATUS 225 

because of the large and often unreliable labor staff which it dispenses 
with. Hand shoveling is sometimes necessary even with modern un- 
loading devices on account of the freezing of coal in the cars. This is 
particularly true of washed coals, and it is not unusual to have an 
entire car load solidly frozen. In this case it has to be picked and 
shoveled by hand, or the unloading tracks must be equipped with 
steam 'pipes and outfits for thawing purposes. A good man is capable 
of shoveling 40 to 50 tons of coal in eight hours when unloading a car, 
provided it is only necessary to shovel the coal overboard. 

134. Bucket Conveyors. — One of the most common methods of auto- 
matically handling the coal from car to bunker is by means of an end- 
less chain of traveling buckets. Many of the largest central stations 
in this country are equipped with such systems. The details of opera- 
tion are best illustrated by a few examples. 

Fig. 134 gives a diagrammatic arrangement of the link-belt, over- 
lapping pivoted bucket carrier, and Fig. 136 illustrates its application 
to a typical boiler plant. Coal is discharged from the railway cars 
into a track hopper and from there delivered by a "feeding apron" 
into a crusher which reduces it to such a size as can be conveniently 
handled by the stokers. It is then discharged into a short bucket 
conveyor, which carries it to the main system of buckets, and it is 
elevated to the proper level and discharged into the overhead bunkers. 
The discharge is effected by special tripping devices which engage the 
buckets and turn them over. The ashes are dumped from the ash pit 
through a series of chutes into the lower run of buckets, by which they 
are elevated and discharged into the ash hopper alongside the coal 
bunkers. From the ash hopper the ashes discharge by gravity directly 
into the railway cars below. The system is operated by means of two 
motors, one driving the crusher and the other the main bucket system. 
The buckets are made of either sheet steel or malleable iron. 

In Fig. 134 the coal is fed to the crusher by the " reciprocating feeder," 
which is usually placed directly under the track hopper. The feeder 
consists of a heavy steel plate mounted on rollers and having a recip- 
rocating movement effected by a crank mechanism from the carrier. 
The amount of coal delivered depends upon the distance the plate moves, 
and this can be varied by changing the throw of the eccentric. The 
number of strokes corresponds to the number of buckets. Any size 
coal can be readily handled. When the distance from track hopper 
to carrier is so great that the reciprocating feeder is not practicable a 
continuous or "belt" feeder is used to supply the crusher with fuel. 
The "equalizing gear" is designed to impart a pulsating motion to the 
driving sprocket wheel which will counteract the natural pulsation to 



226 



STEAM POWER PLANT ENGINEERING 




COAL AND ASH-HANDLING APPARATUS 



227 




228 



STEAM POWER PLANT ENGINEERING 




Fig. 136. Coal and Ash-handling System in the Power House of the South Side 
Elevated Railway Company, Chicago. 



COAL AND ASH-HANDLING APPARATUS 



229 




230 



STEAM POWER PLANT ENGINEERING 



which long pitch chains are subject, producing violent increase of the 
normal strain at frequent intervals. This is accomplished by driving 
the spur wheel with an eccentric pinion, causing the pitch line to describe 
a series of undulations corresponding to the number of sprockets on 
the chain wheel. Figs. 136 and 137 show the general arrangement of 
crusher and " cross conveyor" in the old portion of the South Side 
Elevated Power House, Chicago. 

A coal and ash system similar to the one illustrated in Fig. 136 for a 
plant consisting of eight 350-horse-power boilers will cost in the neigh- 
borhood of $8000, completely installed. This does not include the 
cost of coal and ash bunkers. 




Fig. 138. Driving Mechanism of Hunt Conv< 



The Hunt conveyor, Fig. 138, while usually called a " bucket" con- 
veyor, is in fact a series of cars connected by a chain, each having a 
body hung on pivots and kept in an upright position by gravity. The 
chain is driven by pawls instead of by sprocket wheels. The " buckets" 
are upright in all positions of the chain, consequently the chain can be 
driven in any direction. The change of direction of the chain is accom- 
plished by guiding the carriers over curved tracks. The chain moves 
slowly, and the capacity is governed by the size of the buckets. The 
ordinary size buckets carry two cubic feet of coal and move at a rate 
of fifteen buckets a minute, carrying about 40 tons per hour. Two 
methods of filling the buckets are employed, the " measuring" and the 
"spout filler." In the former each bucket is separately filled with a 



COAL AND ASH-HANDLING APPARATUS 



231 



predetermined amount by a suitable " measuring feeder." In the 
latter the material is spouted in a continuous stream, necessitating the 
use of overlapping buckets to prevent spilling of the material. Fig. 139 
shows an application of the Hunt system to the power plant of the 
Rhode Island Suburban Railway, Providence, R. I, 




Fig. 139. Coal and Ash-handling System, Rhode Island Power House. 



Fig. 140 gives a sectional elevation of the coal and ash-handling 
machinery at the power plant of the Commercial National Bank Build- 
ing, Chicago. Underneath the sidewalk on the Clark Street side of the 
building is a coal-storage bin of 600 tons' capacity, served with a bucket 
conveyor. One leg of the conveyor reaches down to a level below the 
track of the Illinois Tunnel Company. By this arrangement coal can 
be delivered either by cars in the tunnel or by wagons from the street. 
In taking coal from storage a gate at the lower extremity of the hopper 
is opened and the coal filling the buckets is elevated and tripped into 
any one of the screw conveyors leading from bucket conveyor to boiler 
hopper. The ashes are shoveled from the ash pits into cars running in 
a cross tunnel under the boiler floor, and by these cars are transferred to 



232 



STEAM POWER PLANT ENGINEERING 



a dump at one side of the boiler room and discharged into Illinois 
Tunnel Company's cars for removal. 




Pig. 140. 



Bucket and Screw Conveyor at Commercial National Bank Building, Chicago, 
Blinois. 



135. Belt Conveyors. — The Robins belt conveyor, Fig. 141, consists 
essentially of a thick belt of the required width driven by suitable 
pulleys and carried upon idlers so arranged that the belt becomes 
trough-shaped in cross section. In the later designs five pulleys are 
employed instead of three as illustrated in order that the line of con- 
tact may more nearly approach the arc of a circle. The belt is con- 
structed of woven cotton duck covered with a special rubber compound 
on the carrying side. The rubber is thicker at the middle than at the 
edges, since the wear is greatest in a line along the center, but the 
thickness of the belt is uniform throughout its entire width. The 
edges are reenforced with extra plies of duck to increase the tensile 
strength. The idlers are carried by iron or wooden framework, and 
are spaced from 3 to 6 feet between centers on the troughing side, 
according to the width of belt and the weight of the load. On the 
return side these distances range from 8 to 12 feet. High-speed rotary 
brushes with interchangeable steel bristles prevent wet, sticky material 
from clinging to the belt. Automatic tripping devices placed at the 



COAL AND ASH-HANDLING APPARATUS 



233 




Fig. 141. 



Guide Pulleys, Robins Belt 
Conveyor. 



proper points cause the material to be discharged where it is needed. 
The trippers consist essentially of two pulleys, one above and slightly 
in advance of the other, the belt running over the upper and under the 
lower one, the course of the belt resembling the letter S. The material 
is discharged into chutes on the 
first downward turn of the belt. 
The trippers may be movable or 
fixed, single or in series. Movable 
trippers are used when it is de- 
sired to discharge the load evenly 
along the entire length, as, for f~fc|j: 
instance, in a continuous row of 
bins, while fixed trippers are em- 
ployed where the load is to be 
discharged at certain and some- 
what separated points. The movable trippers are made in two 
forms, " hand-driven" and " automatic." In the former they are 
moved from point to point by means of a hand crank. The " automatic " 
tripper is propelled by the conveying belt through the medium of gear- 
ing. It reverses its direction automatically at either end of the run 
and travels back and forth continuously distributing its load. It can 
be stopped, reversed, or made stationary at will. The most notable 
installations of this system are at the Hudson and Manhattan Railway 
Company's power house, Jersey City; L Street Station, Edison Illumi- 
nating Company of Boston, and the South Boston Power Station of 
the Boston Elevated Company. 

136. Elevating Tower, Hand-car Distribution. — Fig. 142 illustrates 
the coal and ash-handling installation at the Aurora and Elgin Inter- 
urban Railroad power house, Batavia, 111. Coal is delivered to the 
plant by railroad cars which dump directly into coal hoppers located 
inside a steel structure running the entire length of the building and 
spanned by two railroad tracks. There are 18 hoppers constructed of 
17-inch brick walls fitted with steel-plate bottoms. Subdividing the 
storage space in this manner makes it possible to carry different grades 
of coal, prevents the spreading of fire, and affords a simple construction 
for the support of the railroad tracks. The basement of the boiler 
room extends underneath the hoppers, and two lines of narrow-gauge 
tracks are embedded in the concrete floor. Turntables at the center 
facilitate the switching of cars to the elevators which rise through the 
boiler room close to the chimney. The cars, of one ton capacity each, 
are of special construction, with roller-bearing axles and a combined 
, ratchet lift and friction dump. The filled cars are pushed from un- 



234 



STEAM POWER PLANT ENGINEERING 



derneath the hoppers to two elevators which lift them to the line of 
tracks supported overhead across the boiler fronts. They are then 
pushed to the hoppers suspended above the boiler setting and the coal 
is dumped. These hoppers have a capacity of six tons each. From 







TRACK TO ELEVATOR 



Fig. 142. 



Coal and Ash-handling System at the Power House of the Aurora and Elgin 
Interurban Railway, Batavia, 111. 



the hoppers the coal is fed to the stoker by an ordinary down spout. 
The ashes fall from the stokers into an ash pit, from which they may be 
discharged into ash cars. The ash cars are elevated to a set of tracks 
running at right angles to the main tracks, and are transferred to ash 
bins located directly over the coal bins. Coal and ashes are weighed 
in the small cars. There are ten boilers in this plant and four men are 



COAL AND ASH-HANDLING APPARATUS 



235 



required to handle the coal and ashes. The entire coal and ash-handling 
system cost about $10,000, and the cost of handling the coal and ashes 
is approximately 4 cents per ton. This does not include wages of 
firemen or water tenders. For a description of recent changes made 
in this plant see Elec. Ry. Jour., Apr. 12, 1911, p. 268. 

137.- Overhead Storage, Bucket Hoist. — Fig. 143 gives a general view 
of the coal-handling plant of the Depot Street power house of the 
Cincinnati Traction Company. This installation is a good example of 
an application of the " overhead storage gravity feed" system to an 
existing plant without interfering in any way with its operation. The 



STORAGE 
1600 TONS 



PIT CAPACITY 
SO TONS 




1-TON SELP 

PILLING 
BUCKET 



Fig. 143. Coal and Ash-handling System at the Depot Street Power House of the 
Cincinnati Traction Company. 

system consists essentially of a receiving pit below the car tracks from 
which the coal is hoisted to a series of overhead bins. The coal storage 
is outside the boiler house in an independent structure. The bins are 
of steel framework with concrete floors, and are sufficiently elevated 
to spout coal easily to the stoker magazine. The total capacity of the 
overhead bins is about 1600 tons. The four bins or receiving pits have 
a Capacity of 50 tons each, or approximately one car load, and are so 
situated that all four may be filled simultaneously without shifting the 
train. The coal-handling apparatus consists of a one-ton self -filling 
bucket operated on a three-motor electric crane running on rails at 
the top of the storage bins. The coal is hoisted from the receiving pit 
through suitable shafts in the bin structure and dumped into the over- 



236 STEAM POWER PLANT ENGINEERING 

head hoppers. The maximum capacity of the hoist is 50 tons per hour. 
The labor required to handle the coal from car to bins is performed by 
one man working five hours per day and an assistant engaged a small 
part of the time to dump cars, clean hoppers, etc. The average daily 
coal consumption is approximately 200 tons. The total cost of the 
equipment was about $18,000 for the bins complete and $4500 for the 
coal-handling crane. The cost of handling the coal and ashes is approxi- 
mately 1.5 cents per ton of coal. Including all charges fixed and operat- 
ing the total cost of handling the coal is about 3.5 cents per ton. This 
does not include wages of firemen or water tenders. 

138. Elevating Tower, Cable-car Distribution. — The coal and ash- 
handling system of the new turbine power plant of the Detroit Edison 
Company is a typical example of a large station equipped with elevat- 
ing tower and cable-car distributers instead of the usual bucket con- 
veyor. The system consists essentially of a lofty steel tower in which 
are housed at various levels a track receiving hopper, crushing rolls 
and feeders, weighing hopper, hoisting apparatus, etc., and a small 
cable railway for delivery to the. bunkers. The railroad coal cars enter 
the tower on an elevated trestle 18 feet above grade, below which is 
a track receiving hopper. A two-ton "tub hoist" is filled with coal 
from the bottom of the receiving hopper and elevated to a 20-ton bin 
at the top, 120 feet above ground level. This bin has a grille bottom 
at one side and under the outlet a heavy duty coal crusher, thus allow- 
ing the fine coal to screen through directly while all the larger lumps 
are automatically delivered to the crusher. From the two bins the 
small cable cars are filled for dumping into the desired bunkers over 
the boiler rooms. The cars are arranged for automatic dumping by 
means of adjustable trips which may be located at any point. The 
entire system has a capacity of from 125 to 150 tons of coal per hour 
and is motor-driven. The ash-handling system consists of brick-lined 
concrete hoppers underneath each pair of stokers which discharge their 
contents by gravity into the small cars operated on the track system 
in the boiler-house basement. 

When handling 600 tons per day of 24 hours the cost of operation is 
approximately 20 cents per ton from coal car to ash car. This includes 
wages of firemen and water tenders. 

139. Hoist and Trolley. — Fig. 144 illustrates a very simple and 
economical method of handling coal and ashes as installed by the 
Jeffrey Manufacturing Company at the power plant of the Scioto 
Traction Company. If the coal car is of the dump type the contents 
are discharged directly into the coal pit from which the coal is re- 
moved by grab bucket and transferred either to the overhead bunker 



COAL AND ASH-HANDLING APPARATUS 



237 




w 



238 



STEAM POWER PLANT ENGINEERING 



or to the storage pile. If the coal car is of the gondola type the coal is 
removed directly from the car by the grab bucket. The bucket is hoisted 
and carried on the trolley into the building over the screen hoppers where 
it discharges its contents; the finer particles fall directly into the 
bunker and the larger lumps are automatically delivered to the crusher. 
The grab bucket will take about 98 per cent of the coal in the car, 
leaving only 2 per cent to be handled by hand. Coal is fed to the 
stokers by means of a traveling electric hopper which receives its supply 
from the overhead bunkers. The present capacity of the plant is 50 
tons per hour taken from the car or pit to stock pile. 




Fig. 145. Diagrammatic Arrangement of the "Vacuum" Ash-handling System. 



140. " Vacuum " Ash Conveyor. — Fig. 145 gives a diagrammatic 
arrangement of a recently patented ash-conveying system depending 
upon the velocity of a column of air for moving the ashes. The system 
is simple in operation and low in first cost. One end of special cast- 
iron header F leads to the ash pits of the various boilers by means of 
branch tubes, and the other end is connected with a sealed separating 



COAL AND ASH-HANDLING APPARATUS 239 

chamber A. Each branch pipe is fitted with simple circular openings 
directly underneath each ash-pit door for admitting ashes and which 
are kept covered except when in operation. Exhauster E creates a 
partial vacuum in chamber A and draws in air at a high velocity from 
the opening in the ends of the branch pipes. Ashes raked into the 
pipes through the openings are caught by the rapidly moving column 
of air' and forced into chamber A. The ashes fall to the bottom and are 
fed into the main ash pit by a slowly revolving ash valve B. Air and 
dust are withdrawn from the top of the separator chamber through 
pipe G and discharged to the stack or to waste. A spray is introduced 
into pipe F to reduce dust. The process is a continuous one and the 
ashes may be completely removed from the ash bin without interfering 
with the operation of the exhauster. In a later construction the ash 
bin and separating chamber are included in one chamber, thus doing 
away with the revolving ash valve and the small motor operating it. 
In this latter design the bin is never completely empty, a certain depth 
of ashes being maintained to seal the bottom at all times. 

At the Armour Glue Works, Chicago, 111., this system is applied to a 
boiler plant of thirteen boilers, aggregating 4800 horse power, and 
cost, completely installed, $5600. As originally installed the separating 
chamber had a volume of about 35 cubic feet and the suction intake was 
placed 58 feet above the ash-pit level. The revolving ash valve made 
about 13 r.p.m., and was driven by a one-horse-power motor. In the 
present installation the separating chamber and motor-operated ash 
valve are dispensed with and the discharge pipes lead directly into the 
main ash bin, which has a capacity of 60,000 pounds of wet ashes 
and is constructed of five-sixteenths-inch sheet iron. The exhauster (a 
30-foot Root blower) has a capacity of about 8000 cubic feet per minute 
at 265 r.p.m., and is driven by a 75-horse-power motor. Under normal 
conditions of operation the motor requires 50 horse power when deliver- 
ing 250 pounds of ash per minute, and the vacuum on the suction side 
of the exhauster is 3.3 inches of mercury. The pipe from the ash bins 
to the separating chamber is 10 inches in diameter and is constructed 
of extra heavy chilled cast-iron pipe. The piping from the separating 
chamber to exhauster and to stack is 22 inches in diameter and is con- 
structed of number 16 and number 20 galvanized iron. The ashes are 
raked by hand from the ash pits to the suction openings of the branch 
pipes, and are handled dry, the dust being taken along with the ashes. 
Elbows are soon worn out by the abrasive action of the ashes, and tees 
are used instead, since the accumulation in the "dead" end receives the 
impact and takes up the wear. The cost of handling the ashes in this 
installation is approximately 7 cents per ton. 



240 



STEAM POWER PLANT ENGINEERING 






W 



adu jsneqxs jjy'ZZ 



^-_ 



8dlJ »Bt>BHYff JJVn 







u o « 



5^a 

5 




2 2 



tai; 



Q 

a 

a 

a 
a 



COAL AND ASH-HANDLING APPARATUS 



241 




242 STEAM POWER PLANT ENGINEERING 

141. Cost of Handling Coal and Ashes. — In large stations where a 
number of men are employed to handle coal and ashes only it is a simple 
matter to divide the cost of handling into the various stages, thus: 

1. Cost of unloading cars or barges. 

2. Cost of conveying coal to bunkers. 

3. Cost of feeding coal to furnace. 

4. Cost of removing ashes. 

These costs are usually expressed in cents or dollars per ton of coal 
burned, or in terms of cents or dollars per horse-power hour or kilo- 
watt hour of main prime mover output. Item number 3 is oftentimes 
included under " boiler-room attendance" and items 1, 3, and 4 under 
"coal and ash handling." Not infrequently all four items are included 
under "attendance." So much depends upon the character of stokers 
and furnace, size of boilers, and the like, that general figures on the cost 
of handling the coal and ashes are of little value unless accompanied by 
a description of the equipment. For the sake of general comparison 
the most satisfactory method of expressing the cost is in dollars per ton 
of coal from coal car to ash car. This includes wages of coal and ash 
passers, repair men, and boiler tenders. In small stations the coal 
and ash handling is done by the boiler tenders, in which case it is 
impracticable to separate the items mentioned above, and the cost is 
ordinarily included under attendance. An average figure for handling 
coal by barrow and shovel is not far from 1.6 cents per ton per yard 
up to the distance of five yards, then about 0.1 cent per ton per yard 
for each additional yard. With automatic conveyors the operating 
cost, not including wages of firemen and water tenders, varies with the 
size of plant and the type of conveyor, and ranges anywhere from a 
fraction of a cent per ton to four or five cents per ton. The larger the 
plant and the greater the amount of coal handled the lower will be the 
cost per ton. In comparing the relative costs of manual and automatic 
handling, fixed charges of at least 15 per cent of the first cost of the 
mechanical equipment should be charged against the latter in addition 
to the cost of operation. In large central stations equipped with stokers 
and conveyors and consuming 200 tons or more of coal in twenty- 
four hours, the cost of handling the coal from coal car to ash car, in- 
cluding wages of firemen and water tenders, will range between 10 cents 
and 18 cents a ton. 

142. Coal Hoppers. — Fig. 148 shows a front and side elevation of 
a typical set of stationary weighing hoppers as applied to the boilers 
of the Quincy Point power plant of the Old Colony Street Railway 
Company, Quincy Point, Mass. Each battery of boilers is provided 



COAL AND ASH-HANDLING APPARATUS 



243 




Fig. 149. Traveling Coal Hoppers. 



244 



STEAM POWER PLANT ENGINEERING 



with an independent set of hoppers. The bottoms of the overhead 
coal bunkers lead into the small hoppers A, A. The operation of any 
single weighing hopper is as follows: Coal is fed from the overhead 
bunkers to weighing hopper H by means of valve V. The weight of 
coal in the weighing hopper is transmitted by a system of levers and 
knife edges to the inclosed scale beam J and noted in the usual way. 
The weighed charge of coal is then admitted to the down spout S by 
means of valves similar to those at V. 

Although separate weighing hoppers for each battery, as illustrated 
in Fig. 148, offer many advantages, they are quite costly and it is not 
unusual to install one or more large weighing hoppers mounted on 
overhead traveling carriages so that one may supply a number of 
boilers (Fig. 149). At the Armour Glue Works, Chicago, the coal 



ri 




Fig. 150. Common Slide Coal Valve. 



Fig. 151. Simplex Coal Valve. 



supply is stored in one large overhead bunker of 1000 tons' capacity. 
A five-ton motor-driven traveling hopper receives its supply from this 
central bunker and delivers it to the various boilers. One man operates 
the traveling hopper, tends to the coal valves, and supplies all boilers 
with coal. 

Weighing hoppers are sometimes made automatic; that is, the open- 
ing and closing of valves, feeding of coal, and recording of weight are 
automatically performed by the weight of the coal itself. The scale is 
set for discharges of a certain weight and continues to discharge this 
amount automatically. In the few plants which are equipped with 
automatic weighing hoppers the capacity of the hopper is approximately 
100 pounds per discharge. These hoppers are necessarily more com- 
plicated and more costly than the ordinary weighing hoppers, and it is 
a question whether the advantages offset the extra first cost and main- 



COAL AND ASH-HANDLING APPARATUS 



245 



tenance charges. A small automatic hopper of 100 pounds discharge 
capacity costs approximately $400 as against $250 for the ordinary 
weighing device. For a description of a coal meter see paragraph 429 

143. Coal Valves. — Figs. 150 
to 154 illustrate the principles of 
a few- well-known coal valves. 
They may be conveniently 
grouped into two classes ac- 
cording to the location of the 
coal pocket: (1) those drawing 
the coal from overhead bunkers 
and (2) those drawing from the 
side of a bin. In the first class 
come the simple slide valve and 
the simplex and duplex rotating 
valve. In the latter are the flap 
valve and the rotating valve. 
They are made in various sizes 




Duplex Coal Valve. 



and designs, but those illustrated are examples 
of the most common types. The simple slide 
valve, Fig. 150, is applicable only to small size 
coal and to small spouts, since coarse or lump 
coal may get in the way and prevent proper 
closing. The simplex valve, Fig. 151, consists 
of a rotating jaw actuated by a lever. There 
are no rubbing surfaces, and the jaws cut 
through the material without jamming. The 
duplex valve, Fig. 152, consists of two rotating 
jaws connected to a common actuating lever. 
The jaws move simultaneously, so that even 
a partially open valve delivers the coal cen- 
trally. When closing the valve the flow is 
gradually stopped by the decreasing width of 
the opening and there is but little resistance 
to the movement of the jaws. The largest 
valve can easily be operated by hand. 

The flap valve, Fig. 153, is the simplest form 
for drawing coal from a side bin. It consists 
merely of an iron flap hinged to the bottom of 
the chute. The valve is lowered to let the coal 
run over its top and is raised to stop the flow. It cannot be clogged or 
get jammed in closing. The flap is raised and lowered by a simple lever. 




Fig. 



153. Common 
Coal Valve. 



Flap' 



246 



STEAM POWER PLANT ENGINEERING 



For very large bins, where the valves are to be opened and closed 
frequently, the "Seaton" valve, Fig. 154, is usually preferred. This 
valve consists of two jaws EE f , and TT' pivoted to suitable framework 
at and actuated by lever A . The valve is shown fully closed. Raising 




Fig. 154. "Seaton" Coal Valve. 



lever A causes the cut-off blade EE' to rotate about and permits the 
coal to flow through the space between the edge of the jaw E and the 
end of the chute. The rate of flow is regulated by the width of this 
opening. The cut-off blade does not reach a stop, hence there is no 
possibility of a lump of coal getting in the way and preventing the 
prompt closing of the valve. 



CHAPTER VII. 

CHIMNEYS. 

144. Chimney Draft. — Draft produced by a chimney depends upon 
so many conditions and involves such a large number of variables that 
empirical methods of proportioning, based upon actual performances, 
are more to be relied upon than theoretical calculations. Draft is 
due to the difference in the weight of the column of hot light gases 
in the stack and that of the cooler and heavier surrounding atmos- 
phere, the latter tending to flow into the base and thereby force the 
lighter gases out the top of the stack. The commonly accepted theory 
of chimney draft is based upon Peclet's hypothesis that the flow through 
the furnace flues and chimney may be represented by the equation 



64.4 \ mj 



in which 

h = the head of fluid producing the flow, feet; 

u = velocity of the gases in the chimney, feet per second; 

G = a coefficient to represent the resistance to the passage of air 

through the coal; 
I = total length of the path of the gases, feet; 
m = area of cross section divided by the perimeter; 
/ = a coefficient depending upon the nature of the surfaces over 

which the gases pass. 

From experiments on chimneys and boilers Peclet gives in con- 
nection with this theory the following values of coefficients G and /: 

G = 12, / = 0.012, 

on the basis of 20 to 24 pounds of coal burned per square foot of grate 
surface per hour. On account of the variation in practice of the factors 
u, f, and G and the difficulty of determining them engineers prefer to 
use the modified formulas given further on. 

The theoretical difference of pressure or intensity of draft may be 
determined as follows: 

247 



248 STEAM POWER PLANT ENGINEERING 

Let H = height of chimney in feet; 

• T = absolute temperature of the freezing point, degrees F. ; 
Ti = absolute temperature of the gases in the chimney; 
T 2 = absolute temperature of the outside air; 
P = average atmospheric pressure = 14.7 pounds per square inch; 
P 2 = observed atmospheric pressure; 

W = weight of a cubic foot of air at 32 degrees F. and pressure P; 
Wi = weight of a cubic foot of chimney gas at 32 degrees F. and 
pressure P. 

Then the weight of a cubic foot of hot gas in the chimney will be 

TTip-J- (51) 

and the weight of a cubic foot of cold air outside will be 

The weight of a column of hot gas H feet high and one foot square 
(assuming uniform density and temperature) will be 

WiffJ-^- (53) 

Similarly, the weight of the cold-air column will be 

WH^.^. (54) 

and the difference in pressure or the intensity of draft will be 

D-**^-™} (55) 

where D is in pounds per square foot. 

By making P = P 2 = 14.7, T = 492, W = 0.0807, TTi = 0.084, 
and Di = pressure in inches of water (Di = 0.192 D), equation (55) 
assumes the familiar form 

By assuming W = Wi = 0.081 and P = 14.7 equation (55) may be 
written 



Di = 0.52#P 2 



(A -A)- < 57 > 



This latter form is ordinarily used, where the atmospheric pressure dif- 
fers considerably from that at sea level, as at high altitudes. Table 41 
gives the density of air and chimney gases at various temperatures. 



CHIMNEYS 



249 



Example: Required the maximum theoretical draft obtainable from 
a chimney 150 feet high, atmospheric pressure 14.7 pounds per square 
inch, temperature of outside air 60 degrees F., temperature of chimney 
gases 550 degrees F. 

Here H = 150, T 2 = 460 + 60 = 520, T x = 460 + 550 = 1010. 

Substituting these values in equation (55), 



Di = 150 



Z7.64 7.95\ 
V 520 1( 



1010/ 



1.02 inches of water, 



which is about 20 per cent greater than the draft actually obtained, 
and represents the maximum possible under the given conditions, 
neglecting the resistance offered by the chimney and the pressure 

TABLE 41. 



DENSITY AND SPECIFIC VOLUME 


OF AIR AND CHIMNEY 


3ASES 


AT 








VARIOUS TEMPERATURES. 








Air. 


Chimney Gases. 


t 


« 


V 


d 


t 


d 


t 


d 


t 


d 





11.581 


.935 


.086353 


200 


.06334 


430 


.04695 


660 


.03730 


5 


11.706 


.945 


.085424 


210 


.06239 


440 


.04643 


670 


.03697 


10 


11.832 


.955 


.084513 


220 


.06147 


450 


.04592 


680 


.03665 


15 


11.931 


.965 


.083623 


230 


.06058 


460 


.04542 


690 


.03633 


20 


12.085 


.976 


.082750 


240 


.05971 


470 


.04493 


700 


.03602 


25 


12.211 


.986 


.081895 


250 


.05887 


480 


.04445 


710 


.03571 


30 


12.337 


.996 


.081058 


260 


.05805 


490 


.04398 


720 


.03540 


32 


12.387 


1.000 


.080728 


270 


.05726 


500 


.04353 


730 


.03511 


35 


12.463 


1.006 


.080238 


280 


.05648 


* 510 


.04308 


740 


.03481 


40 


12.589 


1.016 


.079434 


290 


.05573 


520 


.04264 


750 


.03453 


45 


12.715 


1.026 


.078646 


300 


.05499 


530 


.04221 


760 


.03424 


50 


12.841 


1.037 


.077874 


310 


.05428 


540 


.04178 


770 


.03396 


55 


12.967 


1.047 


.077117 


320 


.05358 


550 


.04137 


780 


.03369 


60 


13.093 


1.057 


.076374 


330 


.05290 


560 


.04096 


790 


.03342 


62 


13.144 


1.061 


.076081 


340 


.05224 


570 


.04056 


800 


.03316 


65 


13.220 


1.067 


.075645 


350 


.05159 


580 


.04017 


900 


.03072 


70 


13.346 


1.077 


.074930 


360 


.05096 


590 


.03979 


1000 


.02861 


75 


13.472 


1.087 


.074229 


370 


.05035 


600 


.03942 


1100 


.02678 


80 


13.598 


1.098 


.073541 


380 


.04975 


610 


.03905 


1200 


.02516 


85 


13.724 


1.108 


.072865 


390 


.04916 


620 


.03869 


1300 


.02373 


90 


13.851 


1.118 


.072201 


400 


.04859 


630 


.03833 


1400 


.02245 


95 


13.976 


1.128 


.071550 


410 


.04803 


640 


.03798 


1500 


.02131 


100 


14.102 


1.138 


.070910 


420 


.04749 


650 


.03764 


1800 


.01848 


110 


14.354 


1.159 


.069665 










2000 


.01698 















d = density, pounds per cubic foot. 
■ t — temperature, degrees F. 

s = specific volume, cubic feet per pound. 

v = comparative volume, volume at 32° = 1. 

Density of chimney gas taken 0.085 pound per cubic foot at 32° F. and 29.92 
inches of mercury. 

(Rankine, " Steam Engine," gives the density at 32° F. as varying from 0.084 to 
0.087.) 



250 



STEAM POWER PLANT ENGINEERING 













TABLE 42. 












THEORETICAL 


DRAFT 


PRESSURE IN INCHES OF WATER. CHIMNEY 








100 FEET HIGH. 1 










Temp. 


Temperature of the 


External Air — Barometer, 14 


.7 Pounds per Square Inch. 2 


in the 
















Chim- 
























ney. 


0° 


10° 


20° 


30° 


40° 


50° 


60° 


70° 


80° 


90° 


100° 


200 


.453 


.419 


.384 


.353 


.321 


.292 


.263 


.234 


.209 


.182 


.157 


220 


.488 


.453 


.419 


.388 


.355 


.326 


.298 


.269 


.244 


.217 


.192 


240 


.520 


.488 


.451 


.421 


.388 


.359 


.330 


.301 


.276 


.250 


.225 


260 


.555 


.528 


.484 


.453 


.420 


.392 


.363 


.334 


.309 


.282 


.257 


280 


.584 


.549 


.515 


.482 


.451 


.422 


.394 


.365 


.340 


.313 


.288 


300 


.611 


.576 


.541 


.511 


.478 


.449 


.420 


.392 


.367 


.340 


.315 


320 


.637 


.603 


.568 


.538 


.505 


.476 


.447 


.419 


.394 


.367 


.342 


340 


.662 


.638 


.593 


.563 


.530 


.501 


.472 


.443 


.419 


.392 


.367 


360 


.687 


.653 


.618 


.588 


.555 


.526 


.497 


.468 


.444 


.417 


.392 


380 


.710 


.676 


.641 


.611 


.578 


.549 


.520 


.492 


.467 


.440 


.415 


400 


.732 


.697 


.662 


.632 


.598 


.570 


.541 


.513 


.488 


.461 


.436 


420 


.753 


.718 


.684 


.653 


.620 


.591 


.563 


.534 


.509 


.482 


.457 


440 


.774 


.739 


.705 


.674 


.641 


.612 


.584 


.555 


.530 


.503 


.478 


460 


.793 


.758 


.724 


.694 


.660 


.632 


.603 


.574 


.549 


.522 


.497 


480 


.810 


.776 


.741 


.710 


.678 


.649 


.620 


.591 


.566 


.540 


.515 


500 


.829 


.791 


.760 


.730 


.697 


.669 


.639 


.610 


.586 


.559 


.534 


550 


.863 


.828 


.795 


.762 


.731 


.700 


.671 


.644 


.618 


.593 


.585 


600 


.908 


.873 


.839 


.807 


.776 


.746 


.717 


.690 


.663 


.638 


.613 



1. For any other height multiply the tabular figure by -^r, where H is the height in feet. 

p 

2. For any other pressure multiply the tabular figure by -. where P is the barometric pres- 
sure in pounds per square inch. 

required to impart velocity to the gases. Table 42 has been computed 
from formula (57) and gives the maximum theoretical draft in a chim- 
ney 100 feet high for different flue-gas temperatures. 

The intensity of draft required to produce best results depends upon 
the kind and condition of fuel, the thickness of fire, character of grate, 
and resistance of the breeching, tubes, baffles, dampers, etc. As stated 
above, the loss of draft in the chimney proper approximates 20 per cent 
of the total, that in the breeching is taken as 0.1 inch per 100 feet of flue, 
and 0.05 inch for each right-angle bend; the loss in the boiler varies 
from 0.3 to 0.6 inch, depending upon the type;* the loss in the furnace 
varies between wide limits, and depends upon the kind of fuel and the 
rate of combustion. The curves in Fig. 155 compiled by the Stirling 
Company and published in their book " Stirling" give the furnace drafts 
necessary to burn various kinds of fuels at different combustion rates, 
and give an idea of the influence of the character of the fuel and the 
rate of combustion. 

* Specific figures may be obtained from the manufacturers. 



CHIMNEYS 



251 







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(M31VM JO 83H0NI) 'lid HSV ONV 30VNUru N33MJL33 Q3Uin2)3U J.JVUQ dO 30H0J 



252 STEAM POWER PLANT ENGINEERING 

Example: Determine the probable draft necessary to burn 30 pounds 
bituminous run of mine per hour per square foot of grate when the out- 
side air is 60 degrees F., the temperature of the chimney gases 550 
degrees, and the flue is 100 feet long, with two right-angle bends. 

The losses will be divided approximately as follows: 

Inch. 

Loss in furnace (from curves in Fig. 155) 0.17 

Loss in boiler (average) 0.40 

Loss in flue, 100 feet at 0.10 per 100 0.10 

Loss in turns, 2 X 0.05 0.10 

0.77 
Since the loss in the chimney alone approximates 20 per cent of the 
total, 0.77 -f- 0.80 = 0.96 will be the maximum pressure difference. 
From equation (56), 

Substituting for the given values of D h T h and T 2 in above equation, 



no* uP M 7 - 95 \ 



from which H = 142, height of stack necessary to produce a draft 
of 0.17 inch in the furnace. 

Table 43 gives the results of a test of a 100-foot unlined steel chimney, 
showing the variation in draft at different points in the stack. 

The curves in Figs. 156 and 157 are taken from Bulletin 21, U. S. Bureau 
of Mines, 1911, and are of interest in illustrating the pressure drops 
throughout the boiler for different conditions of operation. Fig. 156 
shows the pressure drops through the combustion chamber and over the 
fuel bed for a hand-fired Heine boiler when the total drop from ash pit 
to uptake is varied and the resistances to flow of gas are kept constant. 
Fig. 157 shows the pressure drops through the same equipment when 
the total drop is varied and the resistances to flow are kept constant. 
The curves show that the drop through any portion of the path of the 
gases bears a constant ratio to the total drop, provided the resistances 
to the flow of gases remain constant. For further data bearing out 
this fact consult the bulletin referred to. 

Theory of Chimney Draft: National Engineer, Dec, 1911, p. 588, Jan., 1912, 
p. 39; Power, March, 1906, Feb., 1900, p. 12; Engr. U. S., Jan. 15, 1903, May 15, 
1902, p. 313; Trans. A.S.M.E., 11-451, 762, 772, 974, 984; Bulletin No. 21, U. S. 
Bureau of Mines, 1911. 

145. Chimney Formulas. — Rational methods of determining the 
height and area of chimneys being cumbersome and unwieldy and of 
doubtful value for practical use, the various empirical formulas out- 



CHIMNEYS 



253 




.2 .3 A .5 .6 .7 .8 .9 1.0 
Pressure - In.of Water below Atmosphere 

Fig. 156. Pressure Drops through Two Parts of Heine Hand-fired Furnace when Total 
Drop from Ash Pit to Uptake is Varied and Resistances to Flow are Kept Constant. 































































































































































1.0 










































































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over fuel bed to uptake)to total pressure 

drop from ash pit to uptake 














































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bed to total drop from ash pit to uptake 



















































































































































































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Rotation between pressure drop 
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.2 .3 .4 .5 .6 .7 .8 .9 1.0 1.1 

Pressure in Uptake-Inches of "Water below Atmosphere. 



12 



Fig. 157. Pressure Drops through Two Parts of Heine Hand-fired Furnace when the 
Total Drop is Varied and Resistances to Flow are Kept Constant. 



254 



STEAM POWER PLANT ENGINEERING 



lined in Table 44 are quite commonly used. They give good results 
within the limits of the assumptions upon which they are based, but 
otherwise may lead to absurd results, their applicability depending 
largely upon the available data covering the various losses with the 
particular kind, quality, and condition of coal, and conditions of oper- 
ation. Occasionally practical and local considerations fix the height 
of the stack irrespective of theoretical deductions. The logical pro- 
cedure is to determine first the height of chimney necessary to produce the 
draft at the desired maximum rate of combustion and then to proportion 
the area by such formulas as (2), (4), or (5), to suit the quantity of fuel 
to be burned. 

The following heights have been found to give good results in plants 
of moderate size: 

Feet. 

With free-burning bituminous coal 90 

With anthracite, medium and large sizes 120 

With slow-burning bituminous 140 

With anthracite pea 150 

With anthracite buckwheat 175 

With anthracite slack 200 

TABLE 43. 

CHIMNEY DRAFT. 

Test of a 100-Foot Unlined Steel Chimney 3 Feet in Diameter at Massachusetts 
Institute of Technology. (Peabody & Miller, "Steam Boilers," p. 121.) 



Over the grate 

At the bridge wall 

Half-way between bridge and back end 

of boiler 

At the back end of boiler 

In uptake near boiler 

In stack 34 feet above grate 

In stack 51 feet above grate 

In stack 68 feet above grate 

In stack 85 feet above grate 



Draft, Inches 
of Water. 



Maximum. Minimum 



0.24 
0.382 



0.410 
0.354 
0.572 
0.440 
0.334 
0.216 
0.122 



0.218 
0.372 

0.374 
0.334 
0.543 
0.414 
0.312 
0.168 
0.086 



Temperature, Fah- 
renheit. 



Maximum. Minimum 



403 
396 
380 
370 
345 



389 
374 
368 
354 
314 



The chimney serves two 80-horse-power boilers. During test one 
was banked and the combustion at the grate of the working boiler was 
19.8 pounds per square foot of grate surface per hour. Coal burned 
per hour, 590 pounds. 



CHIMNEYS 255 

For plants of 800 horse power or more the height of stack should 
never be less than 150 feet, regardless of the kind of coal used. 

Referring to Table 44, formulas (1), (2), (6), (7), and (9) are based 
upon a fuel consumption of 13 to 15 pounds of anthracite and 22 to 26 
pounds of bituminous coal per square foot of grate area per hour. In 
formulas (3), (4), and (9), the diameter is dependent solely upon the 
quantity of coal burned per hour and the height is determined mainly 
by the rate of combustion per square foot of grate. The results accord 
well with practice. With western coals formula (3) gives results rather 
too large and the constant should be 120 instead of 180. Formula (5) 
is perhaps the most used and has met with much approval. It is based 
on the assumptions that 

1. The draft of the chimney varies as the square root of the height. 

2. The retardation of the ascending gases by friction may be con- 
sidered due to a diminution of the area of the chimney or to a lining of 
the chimney by a layer of gas which has no velocity and the thickness 
of which is assumed to be 2 inches. Thus, for square chimneys, 

E = D> -*£ = A -\VA, (58) 

and for round chimneys, 

E = ^(b*'-*J?\ = A- 0.591 VZ. (59) 

For simplifying calculations the coefficient of VA may be taken as 
0.6 for both square and round chimneys, and the formula becomes 

E = A - 0.6 VZ. (60) 

3. The horse-power capacity varies as the effective area E. 

4. A chimney should be proportioned so as to be capable of giving 
sufficient draft to permit the boiler to develop much more than its 
rated power in case of emergencies or to permit the combustion of 
5 pounds of fuel per rated horse power per hour. 

5. Since the power of the chimney varies directly as the effective 
area E and as the square root of the height H, the formula for horse 
power for a given size of chimney will take the form 

H.P. = CE VW, (61) 

in which C is a constant, found by Mr. Kent to be 3.33, obtained by 
plotting the results from numerous examples in practice. 



256 STEAM POWER PLANT ENGINEERING 

The formula then assumes the form 

H.P. = 3.33 EVH, (62) 



or 



from which 



H.P. = 3.33 (A - 0.6 VA) Vfi, (63) 



Table 45 has been computed from equation 5, Table 44. 

Many engineers simply adopt the following proportions: 

Internal area of chimney at top, one-seventh grate area for bitumi- 
nous coal. 

Internal area of chimney at top, one-ninth grate area for anthracite 
coal. 

Example: Determine the area and diameter of a stack for a 2000- 
horse-power plant to operate under the following conditions: Rated 
load 2000 horse power; maximum overload 40 per cent of rated; flue 
150 feet long, with one right-angle bend; average rate of combustion 
20 pounds of bituminous coal per square foot of grate surface per hour; 
atmospheric temperature 60 degrees F.; flue-gas temperature at over- 
load 600 degrees F. ; coal burned per boiler horse power 4 pounds. 

With modern types of steam engines or turbines an overload of 40 
per cent has little effect on the economy of the prime mover, and the 
boiler efficiency is but slightly reduced, but an additional allowance of 
25 per cent should be made in estimating the overload combustion rate. 

The maximum rate of combustion then will be 



■♦(^-i 



pounds per square foot of grate surface per hour. 

The draft required at the point where the flue enters the, chimney, 
considering the various losses, will be found as follows: 

Inch. 

Furnace (see curves, Fig. 155) 0.3 

Boiler (assumed) 0.4 

Flue, 150 feet at 0.1 inch per 100 feet 0.15 

Turns, 1 at 0.05 0.05 

0.9 

From formula (56), 



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258 



STEAM POWER PLANT ENGINEERING 



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CHIMNEYS 259 

Substituting the following values : 

T 2 = 60 + 460 = 520, T x = 600 + 460 = 1060, 

9 
A = maximum draft = 7^ = 1.12 inch, 
U.o 



V520 1060 



whence the necessary height of stack is 

H = 160 feet (approximately). 

Substituting the value of H in Kent's formula, the effective area is 

found to be 

„ 0.3H.P. 0.3X2000 ,_ _ . , 

E = = — = 7= — = 47.5 square feet, 

VH Vl60 

corresponding to an actual diameter of 93 inches. 

The actual velocity of the gases in a chimney is from 0.25 to 0.35 
of that theoretically possible, the latter being based on the assumption 
that the maximum theoretical draft or pressure difference is available 
for producing velocity. (See paragraph 162.) For examples and cal- 
culations from actual tests see National Engineer, Dec, 1911, p. 588, 
and Jan., 1912, p. 39. 

146. Height of Chimneys for Boilers using Oil Fuel. — Experimental 
data relative to chimneys for boilers using oil fuel are rather meager 
and discordant, but a study of a number of recent installations seems 
to indicate that the area need not exceed 50 per cent of that required 
by the same boiler using bituminous coal. A height 80 to 100 feet 
above the grate usually affords sufficient draft to force the boilers 
50 per cent above rating, but in a number of large installations the 
chimneys have been designed on the coal-burning basis so as to provide 
sufficient capacity in case it proves necessary at a future date to revert 
to the use of coal. See Jour. A.S.M.E., Oct., 1912, p. 1499. 

147. Classification of Chimneys. — Chimneys may be grouped into 
three classes according to the material of construction: 

1. Steel. 

2. Reenforced concrete. 

3. Masonry. 

Steel chimneys have many advantages and are finding much favor 
in large power plants, especially where economy of space warrants 
the erection of the stack over the boiler, in which case the structural 
work of the boiler setting answers for both boiler and chimney. Among 
the advantages are: (1) ease and rapidity of construction; (2) less 
weight for a given internal diameter and height; (3) less surface 



260 



STEAM POWER PLANT ENGINEERING 



exposed to the wind; (4) lower cost; (5) smaller space required; 
(6) slightly higher efficiency if properly calked, for there can be no- 
infiltration of cold air as is likely through the cracks in masonry. The 
chief disadvantage is the cost of keeping the stack well painted to 
prevent rust and the corrosive action of the sulphur in the coal. 
Steel chimneys may be: 

1. Guyed. 

2. Self -sustained. 

148. Guyed Chimneys. — Guyed sheet-iron or steel chimneys or 
stacks held in position by guy wires are employed in small sizes on 
account of their relative cheapness. They seldom exceed 52 inches in 
diameter and 75 feet in height. A heavy foundation is unnecessary, 
and the stack may be supported by the boiler breeching. The small 
short stacks are ordinarily riveted in the shop, ready for erection, 
larger sizes being shipped in sections and riveted at the place of instal- 
lation. The guy wires are usually fastened to an angle iron or band 
at about two-thirds the height, and anchored at a distance from the 
base equal to the height of the band above the ground. 

For very tall stacks two sets of guys are used, from four to six wires 
being fastened to each band, and designed to withstand a wind pressure 
of 30 pounds per square foot of projected area of the stack. Turn- 
buckles are employed to equalize tautness. Table 46 gives the thick- 
ness of material, with approximate cost and weight, of guyed stacks of 
different heights and areas. 

TABLE 46. 

APPROXIMATE WEIGHT AND COST OF GUYED SHEET-STEEL CHIMNEYS. 



Height, Feet. 


Diameter, Inches. 


Thickness of Shell, 
B.W.G. 


Approximate Weight 
per Foot, Pounds. 


40 


18 


16 


13 


45 


20 


16 


14 


45 


22 


14, 16 


20, 15 


50 


24 


14, 16 


22, 16 


50 


26 


14 


23.5 


55 


28 


14 


25 


60 


30 


12, 14 


34, 27 


65 


32 


12, 14 


36, 28 


70 


34 


10, 12 


48, 39 


75 


36 


10, 12 


51, 41 



Approximate cost per pound, 4 cents to 10 cents, including cost of sections 
riveted and punched, ready for assembling, the higher figure referring to the smaller 
stacks. 



CHIMNEYS 261 

149. Self-sustaining Steel Chimneys. — Steel chimneys over 52 inches 
in diameter are usually self-supporting. They may be built with or 
without a brick lining, but the lining is preferred, since it prevents 
radiation and protects the inside from the corrosive action of the flue 
gases. Since the lining plays no part in the strength of the chimney, it 
is made only thick enough to support its own weight, and usually of 
a low-grade fire brick or carefully burned common brick or both. In 
average practice the fire brick extends 20 or 30 feet above the breeching, 
the remainder of the lining being of common brick. In chimneys up 
to 80 inches internal diameter, the upper course is 4J inches thick and 
increases 4§ inches in thickness for each 30 to 40 feet to the bottom. 
In larger chimneys about 8 inches is the minimum thickness. The 
lining is generally set in contact with or close to the shell, though a 
space of from 1 to 2 inches is sometimes left between the brickwork and 
the shell to allow for expansion. This space is occasionally filled with 
sand. 

Self-sustaining stacks may be straight or tapered, and are generally 
made with a flared or bell-shaped base whose diameter and length are 
1J to 2 times the internal diameter of the stack. The base is riveted 
to a heavy cast-iron plate bolted to a concrete foundation of sufficient 
mass to insure stability. 

Fig. 158 gives the details of one of the steel chimneys at the power 
house of the South Side Elevated Railroad, Chicago, 111. 

150. Thickness of Plates. — The sheet is thickest at the bottom, 
decreasing toward the top of the stack. The proper thickness for any 
given section may be determined by treating the shaft as a uniformly 
loaded cantilever, the stresses being expressed by the equation 

in which 

P = the total wind pressure in pounds, 

h = length of the chimney in inches to the center of wind pressure 
(h = L/2 for a cylindrical chimney), 

S = safe stress. A low value of 6000 pounds per square inch for 
single-riveted joints and 8000 for double-riveted joints is 
recommended, for the reason that a tube of such large 
diameter with thin walls will hardly fail by rupture accord- 
ing to the formula, but by flattening and bending. 

- = sectional modulus, 
e 

Di = external diameter of the shell, inches, 

D 2 = internal diameter of the shell, inches. 



262 



STEAM POWER PLANT ENGINEERING 




CHIMNEYS 



263 



(66) 



For chimneys under 7 feet in diameter and 150 feet in height the thick- 
ness of plate should not be less than T \ inch, nor less than f inch for 
larger sizes. 

It is customary to make the courses about 5 feet in height for con- 
venience in erection. 

Table 47 gives the dimensions of self-supporting steel stacks as made 
by the Riter Conley Company of Pittsburg, who use the following 
empirical formula in determining the thickness of the shell, 

c = M 

1 0.8 A 2 ' 
in which 

$i = stress per lineal inch of section considered, 
M = wind moment in inch-pounds, and 
Di = diameter of the shaft in inches. 

Allowing 8000 pounds per square inch as the safe stress for single- 
riveted joints and 10,000 for double-riveted joints, the required thick- 
ness is found by dividing Si by 8000 or 10,000. 

Example : Determine the thickness of plate at a section 150 feet 
from the top of a cylindrical steel stack 12 feet in diameter and 200 
feet high. Horizontal seams to be double riveted. 
The total wind pressure on the section is 

150 X 12 X 25 = 45,000 pounds. 
The moment arm is 
H- X 12 = 900 inches. 

Z>i = 144 inches; S = 8000 pounds per square inch. 



STEEL STACKS. 



TABLE 47. 
SIZES OF RITER CONLEY COMPANY, PITTSBURG. 



Diameter 
of Flue. 


Total 

Height. 


Total 

Weight. 


How Made. 


Ft. In. 


Ft. 


Lb. 




5 6 


165 


67,000 


40 ft. of & in., 45 ft. of \ in., 50 ft. of ^ in., 30 ft. of 

1 in. 
30 ft. of ^ in., 50 ft. of £ in., 50 ft. of & in., 30 ft. of 

f in. 
60 ft. of £ in., 60 ft. of ^ in., 30 ft. of f in. 


7 


160 


79,000 


8 6 


150 


94,000 


10 


200 


150,000 


90 ft. of i in., 60 ft. of ^ in., 50 ft. of f in. 


12 


200 


175,000 


35 ft. of I in., 35 ft. of & in., 35 ft. of ^ in., 35 ft. 
of H in., 35 ft. of f in., 25 ft. of §f in. 


11 6 


225 


232,000 


40 ft. of \ in., 40 ft. of & in., 40 ft. of & in., 40 ft. of 
. H * n -> 40 ft - o f f in., 25 ft. of & in. 


12 


255 


256,000 


75 ft. of I in., 65 ft. of ^ in., 55 ft. of f in., 35 ft. 
of ^ in., 25 ft. of £ in. 



™=*§ /J 



264 STEAM POWER PLANT ENGINEERING 

Substituting these values in equation (65), 

\— -or-)' 

45,000 X 1800 _ ftnnn 3.14 /144^-JV\ 
2 ~ 800 ° X "32" 1 144 / 

D 2 = 143.36. 

Now ' = ^ 2 

= 144 - 143.36 

2 
= 0.32 inch. 

The nearest commercial size lies between nine-thirty-seconds and 
five-sixteenths. 

The Riter Conley formula gives for this section 

« M = 45,000 X 900 

1 0.8 A 2 0.8 X 144 2 
= 2440 pounds, 

St 2440 . Qn _ . , 
t = 8000 = 8000 = - 305inch - 

151. Riveting. — The diameter of rivets should always be greater than 
the thickness of the plate but never less than one-half inch. The 
pitch should be approximately 2\ times the diameter of the rivet, and 
always less than 16 times the thickness of the plate. Single-riveted 
joints are ordinarily used on all sections except the base, where the 
joint should be double riveted with rivets staggered, although in very 
large stacks all horizontal seams are double riveted to give greater 
stiffness to the shaft. 

152. Stability of Steel Chimneys. — The wind being ordinarily the 
only force tending to overturn the stack, and the chimney being rigidly 
bolted to the foundation, a condition of stability requires that 

(W e + W P ) ^ be equal to or greater than P (~ + h\ (67) 

in which 

W c = weight of the chimney in pounds, 
Wf = weight of the foundation, 

P = total wind pressure in pounds, 
D, H, and h, in feet, as indicated in the figure. 



CHIMNEYS 



265 



Expressed graphically: Lay off GP, Fig. 159, equal to the total wind 
pressure in direction and amount and acting at the center of pressure 
of the shaft; lay off GW equal to the 
weight of the stack and foundation; find 
the resultant GR and produce it to 
intersect the base line as at R f ; if R' 
falls within the inner third of the base 
the stack is stable, provided, of course, 
that the chimney is properly designed 
and constructed. Therefore the heavier 
the combined weight of the chimney and 
its foundation the more stable the struc- 
ture. (See also paragraph 157.) 

D in Fig. 159 varies from one-tenth to 
one-fifteenth H, depending upon the char- 
acter of the subsoil. For the ordinary 
concrete foundation, Christie (" Chimney 
Design and Theory," p. 57) gives as an 
average value for D 



D = 



HH 



+ 10. 



26,000 
Steel Chimneys: Elec. Rev., Apr. 7, 1911. 



(68) 




Fig. 159. 



153. Brick Chimneys. — By far the greater number of power-plant 
chimneys are of brick construction and usually of circular section, 
though octagonal, hexagonal, and square sections are quite common. 
The round chimney requires the least weight for stability, and the 
others in the order mentioned. Taking the total wind pressure on the 
flat surface of a square stack as unity, the effective pressure, according 
to Rankine, for the same projected area will be 0.75 for the hexagonal, 
0.6 for the octagonal, and 0.5 for the round. Henry Adams, Industrial 
Engineering, March, 1912, p. 199, states that these figures are not in 
accord with modern experiments, and gives the following multipliers: 
for a round chimney, 0.785; for an octagonal chimney, 0.82. 

Brick chimneys may be divided into two general classes: 

1. Single shell, Fig. 160, and 

2. Double shell, Fig. 162. 

The double shell is the more common and consists of an outer shaft 
of brickwork and an inner core or lining extending part way or through- 
out the entire length of the shaft. 



266 



STEAM POWER PLANT ENGINEERING 



-11-2; ^H 



%- 



£ 



8J8M 



iO^H 







■§ 2<*M^ 



Top of 
Foundation~^jM[ 



10 



l-sHL-l-fl 1 ^ 



TOTAL HEIGHT 

ABOVE FOUNDATION 

200 FT. 




SECTION ON B-B 
Fig. 160. Custodis Radial Brick Chimney. 



CHIMNEYS 



267 



BOB 




^ebb! /Ibdb] 


BBBi 




BOB/ 1 BOB/ 


BOB 




BBB EBB/ 


iBBB/ 




|bbb/ ^_/ 


|bbb| 






1 


AebbI 
IIbbb/ 


|]S2|| 


Fig. 


161. Custodis Radial Brick. 



The single shell is the general construction where carefully burned 
and selected brick not easily affected by the heat are used. As the 
inner core or lining is independent 
of the outer shell and has no part 
in the strength of the chimney, the 
rules for determining the thickness 
of the walls are practically the same 
for both single and double shell. 

154. Thickness of Walls. — The 
thickness of the wall should be such 
as to require minimum weight of 
material for the proper degree of 
stability, due consideration being paid to the practical requirements of 
construction. The thickness does not vary uniformly, but decreases 
from bottom to top by a series of steps or courses as in Fig. 163. In 
general, the thickness at any section should be such that the resultant 
stress of wind and weight of shaft will not put the masonry in tension 
on the windward side or in excessive compression on the leeward side. 

For circular chimneys using common red brick for the outer shell 
the following approximate method gives results in conformity with 
average practice: 

t = 4 + 0.05 d + 0.0005 H, (69) 

where 

t = thickness in inches of the upper course, neglecting ornamenta- 
tion, and should, of course, be made equal to the nearest 
dimension of the brick in use. Ordinary red bricks measure 
8J X 4 X 2, 

d = clear inside diameter at the top, inches. 

H = height of stack, inches. 

Beginning at the top with this thickness, add one-half brick, or 
4 inches, for each 25 or 30 feet from the top downwards, using a batter 
of 1 in 30 to 1 in 36. 

The minimum value of t for stacks built with inside scaffolding 
should be 7 inches for radial brick and 8J inches for common brick, 
as a thinner wall will not support the scaffold. Radial brick for chim- 
neys are made in several sizes, so that the thickness of the walls when 
they are used increases by about 2 inches at the offsets. 

For specially molded radial brick or for circular shells reenforced as 
in Fig. 162 the length of the different courses may be much less than 
stated above. The external form of the top is a matter of appearance, 
and may be designed to suit the taste, but should be protected by a 



268 



STEAM POWER PLANT ENGINEERING 




Fig. 162. Brick Chimney at the Power Plant of the Armour Institute of Technology. 




CHIMNEYS 



269 



cast-iron or tile cap and provided with lightning rods. Ladders for 
reaching the top of the chimney are generally located inside of brick 
stacks and outside of steel ones. 

Professor Lang's rule (Eng. Rec, July 20, 1901, p. 53) for determin- 
ing the length of the different courses is (Fig. 163) 



= c(20£ + 60 1 + 0.1056 G + 2.5^ + 656 tan a -0.007 H 



0.453 



V ~ 18.7), 



(70) 



in which 

h = length of the course under consideration, 
C = constant = 1 for a circular, 0.97 for an 

octagonal, and 0.83 for a square, chimney, 
i = increase in thickness for each succeeding 

section in feet, 
G = weight per cubic foot of brickwork, 
p = wind pressure, pounds per square foot, 
a = angle of the internal batter. 
All other notations as indicated in Fig. 163. 

For chimneys over 100 feet in height he recom- 
mends that 100 be used instead of the actual 
height, since the critical point will be in one of 
the lower sections and not at the base. 

If a value of h is obtained which is not contained 
an even number of times in H, it may be slightly 
increased or decreased so as to effect this result. 

To determine the stresses at any section the 
shaft is treated as a cantilever uniformly loaded 
with a maximum wind pressure of 25 pounds per 
square foot. If the tension on the windward side 
subtracted from the compression leaves a positive 
remainder, the chimney will be stable; if the remainder is negative, the 
masonry will be in tension, which it withstands but feebly. The sum of 
the compressive stresses on the leeward side due to wind pressure and 
weight must be less than the crushing strength of the masonry. The 




practice, however, of 



a fixed value for allowable pressure 



irrespective of the height of the stack gives dimensions that are too 
low for small stacks and too high for large stacks. According to 
Professor Lang, compressive stress on the leeward side in pounds per 
square inch with single chimneys should not exceed 



p = 71+0.65L, 



(71) 



270 STEAM POWER PLANT ENGINEERING 

where 

p = pressure in pounds per square inch, 

L = distance in feet from top of chimney to the section in question. 

With double shell p = 85 + 0.65 L. (72) 

The tension on the windward side should not exceed, 

for single shell: p = (18.5 + 0.056 L), (73) 

for double shell: p = (21.3 + 0.056 L). (74) 

Example: Determine the maximum stress in the outer fibers of the 
brickwork at the base of section 8 of the chimney illustrated in Fig. 160 
when the wind is blowing 100 miles an hour.* Assume the weight of 
the brickwork 120 pounds per cubic foot. 

A wind velocity of 100 miles per hour is estimated to exert a pressure 
of 50 pounds per square foot on a flat surface and approximately 25 
pounds per square foot of projected area on a cylindrical surface. The 
height of the chimney to section 8 is 131.4 feet. The projected area 
as computed from the figure is 1800 square feet. Hence p, the total 
wind pressure, is 1800 X 25 = 45,000 pounds. The volume of brick- 
work above section 9 may be calculated, and is 6150 cubic feet, hence 
the weight W = 6150 X 120 = 738,000 pounds. 

The area of the joint at this section is 75.3 square feet, therefore the 
pressure, due to the weight of the superimposed brickwork is 738,000 
divided by 75.3 = 9800 pounds per square foot. To find the stress 
due to the wind pressure, substitute the proper values in equation (65) : 



Ph = s- e = 0.0983 i r 1 D )S. 



Here 

P = 45,000 as computed above, 

h = 55 feet (found by laying out the section and locating the 
center of gravity), 
D t = 16.2, 
D = 12.9, 

whence 

1 ft 94 _ 1 9 Q4 

45,000 X 55 = 0.0983 ' S, 

from which S = 9907 pounds per square foot. 

* A serious difference of opinion exists as to the effective pressure of wind on 
chimneys of different shapes, but in lieu of accurate experimental data to the con- 
trary the figures given herewith may be used with confidence, since a vast number 
of stacks based upon the figures are successfully withstanding gales of from 60 to 80 
miles an hour. 



CHIMNEYS 271 

The net stress on any part of the section is the resultant of that due 
to the weight of the stack and that caused by the wind, the net stress 
on the windward side being 

9907 — 9800 = 107 pounds per square foot, 

which is evidently a tensile stress and should never exceed the value 
given by formula (73) : 

p= (18.5 + 0.056 L) 
= (18.5 + 0.056 X 131.4) 
= 25.8 pounds per square inch 
= 3715 pounds per square foot. 

The net compressive stress on the leeward side is 9800 + 9907 
= 19,707 pounds per square foot, which should not exceed that given 
by formula (71) : 

p = 71 + 0.65L 
= 71 + 0.65 X 131.4 
= 156.4 pounds per square inch 
= 22,521 pounds per square foot. 

(See also analysis of steel-concrete chimney, paragraph 159.) 

155. Core and Lining. — The core or lining of a brick chimney is 
commonly carried to the top of the shaft, though it sometimes extends 
only part of the distance. The inside diameter is generally uniform, 
the offsets being made on the outside. The core and outer shell should 
be independent to prevent injury due to expansion of the core. The 
rules for the thickness of lining in steel chimneys apply also to brick 
chimneys. The batters for the inner and outer shells should be such 
as to allow at least 2 inches clearance between the two shafts at the 
top, and the top should be protected by an iron ring or by a projecting 
ledge from the outer shell. 

156. Materials for Brick Chimneys. — Brick for the external shaft 
should be hard burned, of high specific gravity, and laid with lime 
mortar strengthened with cement. Lime mortar itself is more resist- 
ant to heat, but hardens slowly and may cause distortion in newly 
erected stacks, and hence should be used only when a long time is 
taken in building. Mortar of cement and sand alone is not to be 
recommended, since it does not resist heat well and is attacked by 
carbon dioxide, particularly in the presence of moisture. A mortar 
consisting of 1 part by volume of cement, 2 of lime, and 6 of sand 
may be used for the upper brickwork, 1, 2 J, and 8 respectively for the 
lower part, and 1,1, and 4 respectively for the cap. The harder the 
brick the more cement is necessary, as lime does not cling so well to 



272 



STEAM POWER PLANT ENGINEERING 



hard, smooth surfaces. The inner core may be constructed of second- 
class fire brick, since the temperature seldom exceeds 600 degrees F. 
Lime mortar is invariably used for the core. 

157. Stability of Brick Chimneys. — When there 
is no wind blowing and the chimney is built sym- 
metrically about a vertical axis the pressure due to 
weight is uniformly distributed over the bearing 
surfaces, and the center of pressure lies in the line 
XX, Fig. 164. But when the wind blows the 
pressure exerted tends to tilt the shaft as a whole 
column in the direction of the 
current, and the resultant pres- 
sure at the windward side of 
the base decreases, until, with 
a sufficiently high velocity of 
wind, it may become zero, in 
which case the center of pres- 
sure moves a distance q towards 
the leeward side of the base. 
As soon as the pressure at A 
becomes zero the joint begins 
to open (assuming no adhesion 
between chimney and base) and 
the shaft is evidently in the 
The distance q through which the center 

For 




x 

i 
i 



X 



(A) 



h 

(B) 



Fig. 164. 

condition of least stability. 

of pressure has moved is called the radius of the statical moment. 

any column it may be shown that 



q=-r- (Rankine, ''Applied Mechanics," p. 229). 
jt\.e 



(75) 



in which 

/ = moment of inertia of the section, 
A = area of the section, 

e = distance from the center of the shaft to the outer edge of the 
joint. 

Thus for circular section, 



D 



For square section, 

For annular circular ring, 

For hollow square, 



Q = 



D 

D 2 + d 2 

SD 
D 2 + d 2 

6Z> ' 



CHIMNEYS 273 

The relationship between weight of shaft and wind pressure for the 
condition of least stability is 

Ph = Wq, (76) 

in which 

P = total wind pressure, pounds, 

h = distance in feet from the base line of the section under con- 
sideration to center of gravity of that section, 
W = weight of shaft in pounds above the assumed base line, 
q = radius of the statical moment 

The condition of least stability for round chimneys requires, there- 
fore, that 

Ph - W^-^. (77) 

For many purposes it is sufficiently accurate to assume D = d, and 
equation (77) becomes 

'Ph = W -r for round chimneys, (78) 

Ph = TF— for square chimneys. (79) 

o 

The rule commonly used in Germany, and which is finding some 
favor with engineers in the United States, gives for the condition of 
least stability 

W (| R + i r) = Ph. (Eng. Rec, July 27, 1901, p. 82.) (80) 

Notations as in Fig. 162, all dimensions in feet. 

This permits of a lighter chimney than equation (77), and the maxi- 
mum wind pressure may be assumed to put the joint Jon the wind- 
ward side in tension or even to permit a slight opening of same. 

A rule of thumb for stability is to make the diameter of the base one- 
tenth of the height for a round chimney; for any other shape to make 
the diameter of the inscribed circle of the base one-tenth of the 
height. 

The factor of stability is the quotient obtained by dividing the value 
of q from formula (76) by that from (75). If less than unity, the 
chimney is in tension at the outer fiber on the windward side, and must 
be redesigned unless the tension is less than that allowed by equation 
(73). Calculations for stability should be made for various sections. 



274 STEAM POWER PLANT ENGINEERING 

Example: Analyze the chimney illustrated in Fig. 160 for stability 
at, say, section 8, the following data referring to the portion above the 
base line of this section. 

From the drawing: 

Projected area of the stack, 1800 square feet. 
Volume of brickwork, 6150 cubic feet. 
Outside diameter of base, 16.2 feet. 
Inside diameter of base, 12.9 feet. 
Center of pressure to base line, 55 feet. 
Total height above base line, 131.4 feet. 
Maximum total wind pressure: 

P = 1800 X 25 = 45,000 pounds. 

Weight of shaft: 

W = 6150 X 120 = 738,000 pounds. 

For stability, according to equation (55), 

D 2 + d 2 



Ph < W 



$D 



Substituting the proper values: 

Ph = 45,000 X 55 = 2,475,000 foot-pounds. 



8 



r - '»•- msn - ^™- 



D 2 + d 2 
While Ph is slightly greater than W — jt-=t — , for practical purposes 

the shaft at this section would be called stable under maximum allow- 
able wind pressure. 

For stability, according to equation (80), 

Ph<W(iR + lr), 

Ph = 2,475,000, as determined above, 

pr(I B + -^) = 738,ooo(H + «f) 

= 4,177,000. 

Ph is therefore considerably less than W (| R + J r), and the con- 
dition imposed in equation (80) is more than fulfilled. 

The Design of Tall Chimneys: Henry Adams, Industrial Engineering, March, 
1912, p. 198. Design of a Brick Chimney: Eng. News, May 9, 1912, p. 866. 



CHIMNEYS 275 

158. Custodis Radial Brick Chimney. — Fig. 160 gives the details of 
a 200 X 10-foot radial brick chimney constructed of special molded 
radial brick, formed to suit the circular and radial lines of each section, 
thus permitting them to be laid with thin, even mortar joints. The 
blocks are much larger than common brick and the number of joints is 
proportionately reduced. They are molded with vertical perforations, 
as shown in Fig. 161, which permits thorough burning, thereby in- 
creasing the density and strength and at the same time reducing the 
weight of the block. In laying, the mortar is worked into the per- 
forations about one-half inch. The first 60 feet above the base are 
octagonal in section, with 36-inch walls, and the balance of circular 
section, with walls tapering gradually from 22 inches to 7 J inches in 
thickness. A radial brick lining extends 60 feet from the base as in- 
dicated. The chimney was designed to furnish draft for a 3500-horse- 
power boiler plant and cost, erected, $8,800. The entire weight of the 
chimney exclusive of foundation is 870 tons. 

Radial brick chimneys without the inner lining are likely to be 
unduly affected by heat. 

The tallest chimney in the world (1912), located at Great Falls, 
Mont., is of the Custodis type, and is used for leading off the gases from 
the smelter plant of the Boston and Montana Consolidated Copper 
and Silver Mining Company. The height above the top of the founda- 
tion is 506 feet, and the internal diameter at the top 50 feet. The 
chimney and foundation cost approximately $200,000. 

Custodis Chimney Details: Eng. Rec, Oct. 1, 1904, p. 385; Power, May, 1900, 
p. 12. 

159. Steel-Concrete Chimneys. — The use of concrete reenforced with 
iron or steel for the construction of chimneys is rapidly increasing. The 
advantages claimed for this class of stack are: 

1. Light weight of the whole structure, being but one-third as great 
as an equivalent common brick chimney. The space occupied is much 
less than with either brick or steel stack, on account of the thinness of 
walls at the base and the absence of any flare or bell. 

2. Total absence of joints, the entire structure including foundation 
being a monolith. 

3. Great resisting power against tension and compression. 

4. Rapidity of construction. May be erected at an average rate of 
six feet per day. 

5. Adaptability of the material to any form. 

This type of chimney being comparatively new, little data concern- 
ing depreciation are available, but some which have been in use ten 
years show little or no deterioration. 



276 



STEAM POWER PLANT ENGINEERING 




Grade a 



Fig. 165. Weber Reenforced Concrete Chimney. 



CHIMNEYS 



277 



Fig. 165 gives the details of a Weber steel-concrete chimney erected 
at Portland, Ore., for the Portland General Electric Company. The 
entire structure, foundation and shaft, is a monolith, 238 feet in total 
height and 12 feet internal diameter, weighing only 889 tons. It 
occupies but 168 square feet of ground space at the grade level. The 
weight not including foundation is 470 tons. The stack was erected 
complete in 58 working days, and cost approximately $13,000. 

The cement used was German Portland mixed with select bank sand 
in proportion of one to three, gravel or crushed stone being used only 
in the foundation below the ground. The mortar was used medium 
dry and tamped in the form around the steel reenforcement. 

The shaft is of the double-shell type, with inner core extending 70 
feet above the grade. The core is but 4 inches in thickness at the 
and the outer shell 8 inches. 



Both inner and outer shell are 
reenforced with vertical T bars, 
lj X lj X A inch, of low-carbon 
Bessemer steel, spaced at the base 
24 inches between centers in the 
inner core and 4 inches in the 
outer shell, and increasing in 
spacing to the top, where the dis- 
tance between the bars is 12 
inches. The horizontal rings are 
1 X 1 X J T's spaced 18 inches 
between centers in the core and 
36 inches in the outer shell. The 
steel bars vary from 16 to 30 feet 
in length, and where they meet 
lengthwise are lapped not less 
than 24 inches. The use of differ- 
ent lengths of steel prevents the 
laps from concentrating in any 
given section. 

The tallest chimney of this type 
(1907) was erected for the Butte 
Reduction Works at Butte, Mont. 
Its height is 350 feet and inside 
diameter 18 feet. Fig. 166. 

The following strain sheet gives the Weber Company's analysis of 
the chimney illustrated in Fig. 165, and is based on a wind pressure of 
50 pounds per square foot. Notations as in Fig. 166. 




278 STEAM POWER PLANT ENGINEERING 

Weights. 

Wf = weight of foundation 

= (^1+^^^)150 

= (30 2 .2 + 3 ° 2 ^ 152 3)l50 

= 523,200 pounds. 
150 = weight per cubic foot of concrete. 
We — earth weight on foundation 

= I k 2 h 6 ~ (volume of foundation) j 100 

= (7200 cubic feet - 3995 cubic feet) 100 

= 320,500 pounds. 
100 = weight per cubic foot of earth. 
W = weight of shaft 

= {Ai Qu + h) + A 2 (h + h) + A 3 h 5 } 150 

= 538.5 (72 + 3) + 13 (72 + 3) + 19.75 • 158} 150 

= 934,950 pounds. 
W t = total weight 

= W f + We + W = 1,778,650 pounds (889 tons). 

Section at Grade G r . 

I. Wi = weight of outer shell and single shell above section 
= (A Ju + A s h) 150 
= (28.5 • 72 + 19.75 - 158) 150 
= 775,806 pounds. 
II. r = radius of statical moment 



-f[>+(#J1 



14.66 



8 

= 3.35 feet. 
III. P = wind pressure on chimney 
n , 50 50 

= 14.66 X 72 X 25 + 13 X 158 X 25 
= 77,738 pounds. 
M = wind moment on section 



50 A 4 , /~ 7 50\ /. , h s 



- 14.66 X 72 X |° X 3 +(l3 X 158 X f ) 72 + J« 

= 8,703,818 foot-pounds. 



CHIMNEYS 279 

IV. N = statical moment 

= 3.35 X 775,806 

= 2,598,950 foot-pounds. 

V. B = bending moment 
= M -N 

= 8,703,818 - 2,598,950 
= 6,104,868 foot-pounds. 

VI. - = section modulus 

'-- .m 

-0.0982 ("- 66X1 ^ 3 - 33X12 ) 

= 169,703. 
VII. z = tension per square inch sectional area 

= 12B + - 

e 

= 12 X 6,104,868 ^ 169,703 
= 432.5 pounds. 

VIII. Z = total tension 
= 144 Ai« 

= 144 X 28.5 X 432.5 
= 1,825,015 pounds. 

IX. s = area steel required 

= — • (a = sectional strain on steel) 

= 16,000 pounds per square inch 
= 114.2 square inches. 

X. K = number of bars 

= — • (x = 0.45 square inch = area of one bar) 
x 

= 252 bars. 

For Stability. 

XI. L = length of one side of base. 

8,703,818 
1,778,650 X 
= 29.4 feet. 



280 STEAM POWER PLANT ENGINEERING 



Section 42' 0" above Gkade. 


I. 


Wi = 596,250 pounds. 


II. 


r = 3.35 feet. 


III. 


P = 62,295 pounds. M = 5,761,325 foot-pounds. 


IV. 


N = 1,997,438 foot-pounds. 


V. 


B = 3,763,889 foot-pounds. 


VI. 


- = 169,703. 
e 


VII. 


z = 222 pounds. 


VIII. 


Z = 911,088 pounds, 


IX. 


s = 57 square inches. 


X. 


K = 127 bars. 


Section 


at Offset. 


I. 


Wi = 468,000 pounds. 


II. 


r = 3 feet. 


III. 


P = 51,350 pounds. M = 4,056,650 foot-pounds. 


IV. 


N = 1,404,000 foot-pounds. 


V. 


B = 2,652,650 foot-pounds. 


VI. 


- = 102,041. 

e 



VII. z = 311 pounds. 

VIII. Z = 786,000 pounds. 
IX. s = 55.5 square inches. 

X. K = 123 bars. 

Section 50' 0" from Top. 

I. W x = 148,125 pounds. 
II. r = 3 feet. 

III. P = 16,250 pounds. M = 406,250 foot-pounds. 

IV. N = 444,365 foot-pounds. 

Since the statical moment N is greater than the wind moment M, 
there is no bending moment B, so no steel is required, the chimney 
above this section standing of its own weight. However, thirty-two 
bars are continued to the top. 

Analysis of Reenforced Concrete Chimneys: Eng. Rec, Apr. 8, 1911; Jan. 13, 
1912. 

160. Breeching. — The area of the flue or breeching leading from the 
boilers to the chimney is generally made equal to or a little larger than 
the internal area of the chimney, 20 per cent greater being an average 
figure. The flue may be carried over the boilers or back of the setting 
or even under the fire-room floor, but in any case should be as short as 



CHIMNEYS 281 

possible and free from abrupt turns. Short right-angled turns reduce 
the draft approximately 0.05 inch for each turn, and a convenient rule 
is to allow 0.1 inch loss for each 100 feet of flue if of circular cross sec- 
tion and constructed of steel, and double this amount for brick flues 
of square section. The cross section of the flue need not be the same 
throughout its entire length, but may be tapered and proportioned to 
the number of boilers. Where two flues enter the stack on opposite 
sides, a diaphragm is inserted as indicated in Fig. 160. Flues should 
be covered with heat-insulating material. 

161. Chimney Foundations. — On account of the concentration of 
weight on a small area the foundation of a chimney should be carefully 
designed. In most cities the building laws limit the maximum loads 
allowed for various soils and materials, and although they vary con- 
siderably the average is approximately as follows: 

Material. Safe Load, Lb. per Sq. Ft. 

Hard-burned brick masonry, cement mortar, 1 to 2 20,000-30,000 

Hard-burned brick masonry, cement mortar, 1 to 4 18,000-24,000 

Hard-burned brick masonry, lime mortar 10,000-16,000 

Concrete, 1 to 8 8,000-10,000 

Kind of Soil. Safe Load, Tons per Sq. Ft. 

Quicksands and marshy soils 0.5 

Soft wet clay 1.0 

Clay and sand 15 feet or more in thickness 1.5 

Pure clay 15 feet or more in thickness 2.0 

Pure dry sand 15 feet or more in thickness 2.0 

Firm dry loam or clay 3.0- 4.0 

Gravel well packed and confined 6.0- 8.0 

Rock broken but well compacted 10.0-15.0 

Solid bed rock Up to \ of its ultimate crushing strength. 

Tons per Pile. 

Piles in made ground 2.0 

Piles driven to rock or hardpan 25.0 

Chimney foundations as a rule are constructed of concrete except 
where the low sustaining nature of the soil necessitates the use of piles 
or a grillage of timber or steel. For masonry chimneys the foundation 
is designed to give the necessary support to the shaft without particular 
reference to its mass or distribution, as the shape of the foundation has 
virtually no effect on its stability as a column. In steel and reenforced 
concrete chimneys the shape and weight of the foundation are a function 
of the desired factor of stability, since the shaft is securely anchored 
to the foundation and the two form practically one mass. The founda- 
tion should be designed to fulfill the conditions in formula (68) in 
addition to the requirements for mere support. 



282 



STEAM POWER PLANT ENGINEERING 



Table 48 gives the least diameter and depth of foundation for steel 
chimneys of various diameters and heights. 

162. Chimney Efficiencies. — The chimney as a mover of air has a 
very low thermodynamic efficiency. Compared with that of a fan its 
performance is very poor, and mechanical-draft concerns sometimes 
use this as an argument. 

Example: A chimney 200 feet high and 10 feet in diameter furnishes 
draft for a battery of boilers rated at 3500 horse power. Average out- 
side temperature 60 degrees F. ; temperature of flue gases 500 degrees F. ; 
calorific value of the fuel 14,000 B.t.u. per pound. Compare the ther- 
mal efficiency of the chimney as a mover of air with that of a forced- 
draft apparatus of equivalent capacity. 

TABLE 48. 

SIZES OF FOUNDATION FOR STEEL CHIMNEYS. 



Diameter, Feet. 


Height, Feet. 


Least Diameter of 
Foundation. 


Least Depth of 
Foundation. 


3 


100 


15' 9" 


6' 0" 


4 


100 


16' 4" 


6' 0" 


4 


125 


18' 5" 


7' 0" 


5 


150 


20' 4" 


9' 0" 


5 


200 


23' 8" 


10' 0" 


6 


150 


21' 10" 


8' 0" 


6 


200 


25' 0" 


10' 0" 


7 


150 


22 r r 


9' 0" 


7 


250 


29' 8" 


12' 0" 


9 


150 


23' 8" 


10' 0" 


9 


275 


33' 6" 


12' 0" 


11 


250 


24' 8" 


10' 0* 


11 


350 


36' 0" 


14' 0" 



From Table 42 we find tHat a chimney 200 feet high, with tempera- 
tures as stated above, will furnish a theoretical draft of 1.27 inches, 
equivalent to a pressure of 6.6 pounds per square foot. Neglecting 
friction the height H of a column of external air which would produce 
this pressure is fd 1 - d\ , /01 , 

= V — d — / ' 
in which 

h = height of the chimney in feet, 
d = density of the hot gases in the stack, 
di = density of the outside air. 
Substitute in (59) 

di = 0.0763, d = 0.0435, and h = 200. 
= / 0.0763 - 0.0435 N 
V 0.0763 / 
= 85.9 feet. 



200 



CHIMNEYS 283 

The theoretical velocity of the air entering the base of the chimney 
under this head is 

v= VYgH 
= V2 X 23.2 X 85.9 
= 74.5 feet per second. 

The weight of the gas escaping per second 

= 74.5 X area of the stack X 0.0763 
= 446 pounds. 

The displacement of this volume of gas is the result of heating it from 
60 to 500 degrees F. Taking the specific heat of the gas as 0.2375, the 
heat necessary to displace 446 pounds per second is 

Heat required = 446 X 0.2375 X (500 - 60) 
= 46,500 B.t.u. per second. 

The work actually performed is that of overcoming a total resistance 
of 6.6 X 78.5 = 518 pounds (78.5 = internal area of the chimney) 
through a space of 74.5 feet; i.e., 

Work done = 74.5 X 518 = 38,591 foot-pounds per second 

= 49.7 B.t.u. per second. 

49.7 
Efficiency = , fi ' n = 0.00107, or about T V of 1 per cent. 

If a fan be substituted for the chimney and we allow say 8 per cent 
for the efficiency of engine and boiler, 40 per cent for the fan, and 25 per 
cent for friction, the combined efficiency will be 

0.08 X 0.40 X 0.75 = 0.024, or 2.4 per cent. 

0.024 
The fan then will be ' =22.4 times more efficient than the 

chimney as a mover of air. 

163. Cost of Chimneys. — Christie ("Chimney Design and Theory ") 
gives the following costs of chimneys 150 feet high and 8 feet internal 
diameter: 

Common red brick approximate cost $8,500.00 

Radial brick do do 6,800.00 

Steel, self-supporting, full lined do do 8,300.00 

Steel, self-supporting, half lined do do 7,800.00 

Steel, self-supporting, unlined do do 5,820.00 

Steel, guyed do do 4,000.00 



284 



STEAM POWER PLANT ENGINEERING 



The following approximate costs of various sizes of a well-known 
radial brick chimney give an idea of the variation in cost due to in- 
crease in diameter and height: 



Size of Chimney. 




Size of Chimney. 








Cost. 






Cost. 


Height. 


Diameter. 


Height. 


Diameter. 


Feet. 


Feet. 




Feet. 


Feet. 




75 


4 


$1,350.00 


175 


8 


$7,050.00 


75 


6 


1,950.00 


175 


10 


7,925.00 


75 


8 


2,650.00 


175 


12 


8,950.00 


75 


10 


3,725.00 


175 


14 


9,725.00 


125 


6 


3,500.00 


200 


8 


9,250.00 


125 


8 


4,250.00 


200 


10 


10,500.00 


125 


10 


4,675.00 


200 


12 


11,100.00 


125 


12 


5,125.00 


200 


14 


12,500.00 ' 


150 


8 


6,150.00 


250 


10 


16,500.00 


150 


10 


7,125.00 


250 


12 


18,250.00 


150 


12 


7,750.00 


250 


14 


21,500.00 


150 


14 


8,275.00 


250 


16 


24,250.00 





TABLE 49. 

PROPORTIONS OF CHIMNEYS FOR FACTORY STEAM BOILERS, 
FROM PRACTICE. (Hutton.) 



COLLECTED 





Internal Dimensions. 




Thickness of Walls. 


Height of 






Ratio of 






Chimney 
above the 






Bottom to 
Top. 










Thickness 




Ground in 






Internal 


at Base in 


Thickness 


Feet. 


Ground Line. 


Size of Top. 


Area. 


Inches at 
Ground 

Line. 


at the Top 
in Inches. 


40 


2' 6" 


1' 9" sq. 


2.04 


18 


9 


60 


2' 11" 


2' 0" sq. 


2.12 


18 


9 


70 


3' 4" 


2' r sq. 


2.13 


23 


9 


80 


3' 8" 


2' 6" sq. 


2.18 


28 


9 


90 


4' 0" 


2' 9" sq. 


2.27 


28 


9 


100 


4' 8" 


3' 0" diam. 


2.40 


28 


9 


110 


4' 10" 


3' 3" diam. 


2.33 


28 


9 


120 


5' 6" 


3' 6" diam. 


2.40 


28 


9 


135 


6' 0* 


4' 0" diam. 


2.30 


28 


9 


150 


4' 6" 


3' 0" diam. 


2.25 


28 


14 


155 


6' 0* 


4' 6" diam. 


1.78 


56 


14 


160 


9' 0* . 


5' 0" sq. 


3.24 


36 


14 


170 


7' 6" 


5' 0" diam. 


2.25 


36 


14 


180 


6' 4" 


4' 6" diam. 


2.00 


54 


14 


200 


5' 3" 


3' 6" diam. 


2.28 


36 


14 


225 


16' 0* 


6' 6" sq. 


4.00 


36 


14 


250 


19' 0" 


13' 0* diam. 


2.13 


40 


14 


300 


14' 0" 


9' 0" diam. 


2.42 


48 


14 


450 


21' 6" 


10' 2" diam. 


4.35 


59 


14 



CHAPTER VIII. 

MECHANICAL DRAFT. 

164. General. — The intensity of natural draft in a chimney depends 
mainly upon the height of the stack and the temperature of the chim- 
ney gases, and the chimney should be designed to meet the maximum 
requirements, permitting the damper to be partly shut at times. There 
is usually no practicable means of increasing natural draft per se after 
the maximum has been reached. Again, chimney draft is peculiarly 
susceptible to atmospheric influence and may be seriously impaired 
by adverse winds and air currents. Notwithstanding these apparent 
limitations, by far the greater number of steam power plants depend 
upon chimneys for draft because of the disposition of the waste gases. 
In many cases artificial draft has a great advantage and under certain 
conditions is indispensable; it is very flexible and readily adjusted to 
effect various rates of combustion, irrespective of climatic influences, 
and permits any degree of overload without undue expenditure of 
energy. 

Artificial draft may be broadly classified under two heads : 

1. The vacuum or induced draft; and 

2. The plenum or forced-draft method. 

In the former a partial vacuum is produced above the fire by suitable 
apparatus, and the effect is substantially that of natural draft. 

In the forced-draft system pressure is produced in the ash pit, the 
air being forced through the grate. 

In both systems the artificial draft is usually produced by either: 

1. Steam jets; or 

2. Centrifugal fans or exhausters. 

165. Steam Jets. — Fig. 167 shows an application of a ring jet to 
the base of a stack. The apparatus is very simple, inexpensive, and 
easily applied. It consists essentially of a ring or a series of concen- 
tric rings of 1-inch or lj-inch pipe, perforated on the upper side with 
Tt> or J-inch holes, and placed in the base of the stack, so that the jets 
are discharged upward, thus creating a draft independent of the temper- 
ature of the flue gases. The steam connection to the jet is generally 
made direct to the boiler and not to the steam main, though the jet 
is often produced by exhaust steam. 

285 



286 



STEAM POWER PLANT ENGINEERING 



Fig. 168 illustrates a Bloomsburg jet, which involves to some extent 
the principle of the ejector. 






Fig. 167. Ring Steam Jet. 



Fig. 168. Bloomsburg Jet. 



The increase in draft produced by these devices as ordinarily in- 
stalled is not great, although in locomotive practice where the entire 
exhaust is discharged up the stack an intense draft is obtained. 




Fig. 169. McClaves Argand Blower. 

Fig. 169 shows the application of a " McClaves argand blower/ 7 
The steam is discharged below the grate through a perforated hollow 
ring, as indicated, drawing the air through the funnel by inspiration. 
This creates a powerful draft by forming an air pressure in the ash pit, 
and is an especially useful system of forcing fires for boilers which need 
forcing for short periods only. 



MECHANICAL DRAFT 



287 



Steam jets, as ordinarily installed in small plants, are very uneconom- 
ical, since a large amount of steam is required to produce good results. 
Table 50, based on experiments at the New York Navy Yard, to deter- 
mine the best form of steam jet for producing draft in launch boilers, 

TABLE 50. 

RESULTS OF EXPERIMENTS UPON STEAM JETS AT NEW YORK NAVY YARD.* 





Pounds of Water Evaporated per Hour. 


Index of Jet. 


A 


B 


C 


D 


E 


In boiler making steam 

In boiler supplying jets 

Per cent of steam used 
bv iet 


463.8 
97.5 

21.2 


580.0 
120 

20.7 


361.25 
30 

8.3 


528.5 
63.2 

12.0 


545.00 
76.25 

19.0 







* Annual Report of the Chief of the Bureau of Steam Engineering, U. S. Navy, 1890. 

shows steam consumptions of from 8.3 to 21.2 per cent of the total 
steam made. Table 51 gives the steam consumption of a number of 
types of steam jet blowers as determined by A. J. Whitman. The best 
performance is 4.6 per cent and the poorest 11.1 per cent of the total 

TABLE 51. 

CONSUMPTION OF STEAM BLASTS COMPARED, t 



Coal. 


Name of Blower. 


Per Cent of Air 

Openings in 

Grate. 


Pounds of Dry 

Coal burned per 

Hour per Square 

Foot of Grate. 


Per Cent of Total 
Steam Generated 

in the Boilers 

that is required 

to operate the 

Steam Blasts. 


Rice 


Young 


11 

11 

7 

11 

11 

26 

11 

11 

7 

7 


25.8 
17.9 
27.0 
27.3 
16.7 
31.4 
16.4 
26.1 
32.5 
45.4 


11.1 
7.0 

10 8 


Do 

Do 


do 

Wilkinson 

Young 

. . . .do 


Buckwheat 

Do 


10.8 
4 6 


Do 


.do 


8 9 


Do 

Do 


McClave 

....do 


6.7 
9 3 


Do 


Wilkinson 

....do 


7.8 


Do 


10.2 











t Trans. A.S.M.E., Vol. XVII. — See Whitham. 

boiler steam generated. Steam jets below the grate are said to prevent 
clinkers from forming where fine anthracite coals are used, and thus 
to assist in keeping the fire free and open. They also assist in the 
economical combustion of certain low-grade fuels. See paragraph 102 
for the influence of steam jets in effecting smokeless combustion. 



288 



STEAM POWER PLANT ENGINEERING 



The curves in Fig. 170 are of interest in showing the intensity of 
draft created by steam jets in the modern locomotive and the influence 
of the draft on the capacity of the boiler and the air supply. These 
curves are taken from Bulletin 21, U. S. Bureau of Mines, 1911. 



oh 

o 
513 

on 



2 
o 
W 

© 
P. 

s ^ 



100 



2 



fe 



"eg o 

O p. 



20 



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VUL' 


















































































































































































































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3 




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11 


12 



1000 



900 



800 



,uu„ 



600 "2 



o 
500 > 

ft 

400 | 

o 

300 I 
o 

w 

200 



100 



Fig. 170. 



Total Pressure Drop-Inches of Water. 
Influence of Draft on the Performance of a Locomotive Boiler. 



In large modern central stations where boiler overloads of from 150 
to 250 per cent above rating are desirable, steam jets and mechanical 
blowing and stoking appliances use but a nominal percentage of the 
steam generated. The results in Table 52, taken from the tests of the 

TABLE 52. 

STEAM CONSUMPTION OF DRAFT APPLIANCES AND STOKER ENGINES, 2365 H.P. 

STIRLING BOILER. 
(Delray Station, Detroit Edison Co.) 



RONEY STOKER. 



No. of 


Per Cent 
of Rating. 


Dry Coal 

per Sq. Ft. 

G. S. per 

Hr. 


Steam Consumption, Per Cent 
of Total Generated. 


Draft, Inches of Water. 


Test. 


Stoker 
Engines. 


Steam 
Jets. 


Total. 


Below 
Dampers. 


In 

Furnace. 


Ash 
Pit. 


5 

4 

18 


94 
152 
195.7 


14.81 

25.97 
33.60 


0.19 
0.15 
0.13 


1.56 
1.43 
1.19 


1.75 
1.58 
1.32 


0.16 
0.55 
1.11 


0.24 
0.22 
0.33 


0.10 
0.02 
0.05 



MECHANICAL DRAFT 



289 



TABLE 52 — Concluded. 



TAYLOR STOKER. 





Per Cent 
of Rating. 


Drv Coal per 

Sq. Ft. G. S. 

per Hr. 


Steam 

Consumption 
of Stoker 

Engines and 
Turbine 
Blower. 


Draft, Inches of Water. 


No. of 
Test. 


At Blast in 
Tuyeres. 


Suction Below 

Boiler 

Dampers. 


Suction in 
Ash Pit. 


10 

9 

11 


92.9 
162.8 
211.0 


16.43 
29.23 
38.75 


2.63 
2.87 
3.41 


0.67 
1.73 
2.53 


0.20 
0.53 
0.84 


0.15 

0.06 
0.02 



All of the steam exhausted from the Taylor equipment may be returned to the feed-water heater, 
whereas only that exhausted from the engines in the Roney equipment may be used in this manner, hence 
the net heat used is approximately the same in both cases. 

For application of steam jets to mechanical stokers see Chapter IV. 

large Stirling boilers at the Delray Station of the Detroit Edison Com- 
pany, show what may be expected from installations of this class 
(Jour. A.S.M.E., Nov., 1911). 

166. Fan Draft. — Fig. 171 shows a typical installation of a centrif- 
ugal fan on the forced-draft or plenum principle, the fan creating a 
pressure in the ash pit and forcing air through the fuel. The most 
approved method is to pass the air through the bridge wall, thence 
toward the front of the grate, though it may enter through an under- 
ground duct or through the side of the setting. Forced draft is usually 
adopted in old plants where increased demands for power require that 
the boilers be forced far above their rating to save the heavy expense 
of new boilers, or in plants burning refuse, anthracite culm or screen- 
ings, which require an intense draft for efficient combustion. Forced 
draft is also well adapted for underfeed stokers of the retort type, 
hollow blast grates, and the closed fire hole system. The air supply 
may be taken from an air chamber built around the breeching, thereby 
supplying the heated air to the fan and effecting a lower temperature 
in the breeching and a higher temperature in the furnace. The ob- 
jection is sometimes raised against forced draft that the gases tend to 
pass outward through the fire door when the fire is cleaned or re- 
plenished, since the pressure in the furnace is greater than atmospheric. 
This objection may usually be overcome by suitable dampers in the 
blast pipe which are closed on opening the fire doors. With a boiler 
plant of 1000 horse power or more the cost of a forced-draft fan, engine, 
and stack will approximate from 20 to 30 per cent of the outlay for an 
equivalent brick chimney. The power consumption will depend upon 
the character and efficiency of the motor or engine and will range from 
1 to 5 per cent of the total capacity. 



290 



STEAM POWER PLANT ENGINEERING 



Induced draft as illustrated in Fig. 172 is perhaps the most com- 
mon substitute for natural draft and is extensively used in street rail- 
way and lighting plants which have high peak loads, being ordinarily 
installed in connection with fuel economizers. The suction side of the 
fan is connected with the uptake or breeching of the boiler or bat- 
teries of boilers and the products of combustion are usually exhausted 
through a stub stack. The illustration shows a typical installation in 
which two fans of the duplex type are placed above the boiler setting. 
The fan ducts are generally designed with a by-pass direct to the stack 
to be used in case of accident or when mechanical draft is not required. 




Fig. 171. Typical Forced-draft System. 

Since the fan handles hot gases it must, under the ordinary con- 
ditions of practice, have a capacity approximately double that of a 
forced-draft fan delivering cold air, but the gases being of lower density 
the power required per cubic foot moved is less. 

With forced draft about 300 cubic feet of air are required per pound 
of coal; with induced draft the fan must handle twice this volume if 
the gases are exhausted at 500 degrees F. or 450 cubic feet if exhausted 
at 300 degrees F., a temperature to be expected in connection with 
economizers. 

The advantages of induced draft over forced draft are very pro- 
nounced. The pressure in the furnace is less than atmospheric, there- 
fore it is not necessary to shut off the draft in cleaning fires or ash pit, 
and the fire burns more evenly over the entire grate area, since the 
draft pressures are ordinarily less than with forced draft. An induced- 
draft plant costs considerably more than forced draft on account of 
the larger fan required, but the operating expenses are but little greater. 



MECHANICAL DRAFT 



291 



With a boiler plant of 1000 horse power or more the cost of a single 
induced-draft fan, engine, stack, etc., will approximate from 40 to 50 per 
cent of the outlay required for a brick chimney of equivalent capacity, 
and the double-fan outfit will approximate from 50 to 60 per cent. 
The double-fan system is particularly adapted to plants which operate 
continuously and where even a temporary break-down is a serious in- 
convenience. 




Fig. 172. Typical Induced-draft System. 



Turbo-undergrate draft blowers, installed in each setting, are finding 
favor with many engineers because of the low cost of installation. 

They consist essentially of small impulse steam turbines direct con- 
nected to specially designed propeller fans set in the side walls of the 
setting by means of wall thimbles. The fan discharges below the 
grate, and may be automatically controlled by damper regulation. 
The turbine exhaust may be discharged into the ash pit to prevent 
clinkers, or it may be used in the feed-water or other heating devices. 
They are more economical in heat consumption than the ordinary jet 
device. 



292 



STEAM POWER PLANT ENGINEERING 



167. Performance of Fans. — The first satisfactory theory of centrif- 
ugal fans was promulgated by Daniel Murgue in 1872. He proved 
that theoretically the maximum pressure created by a perfect fan is 
equivalent to twice the head which would produce a velocity equal to 
that of the periphery. Thus 

,1.2 

H = -, (82) 

in which 

H = maximum difference in pressure in feet of air, 
u = peripheral velocity in feet per second, and 
g = acceleration of gravity 32.2. 

A and B, Fig. 173, represent Pitot tubes inserted in the discharge 
pipe of a centrifugal blower, A being bent to face the current, while B 
is at right angles to it. A receives the full impact of the stream, and 



imic Opening 



/-> 



(A) 



Static ( B ) 
Opening 



Orifice Closed 



tion or eddy currents, 



Fig. 173. 

the manometer indicates the total pressure, static and velocity, while 
B registers the static pressure only. With the discharge orifice closed, 
as in Fig. 173, the velocity becomes zero, and the water depression in 
both manometers will be the same, due to the static pressure, which 
according to Murgue's theory, will be a maximum and, ignoring fric- 

u 2 

9 ' 

Example: Determine the maximum pressure, in inches of water, 
which a perfect fan would exert with discharge orifice closed; diameter 
of fan 6 feet; r.p.m. 318. 
The peripheral velocity is 

u = 2irrn = 6.28 X 3 X 318 = 6000 feet per minute 

= 100 feet per second. 
Substituting in Murgue's formula, 

u = 100 and g = 32.2, 

100 2 
H = W2 = 31 ° feet ' 



MECHANICAL DRAFT 



293 



i.e., the pressure created by the fan would be equivalent to the weight 
of a column of air 310 feet high, or, assuming an air temperature of 
75 degrees F., an equivalent head in inches of water of 
310 X 0.074495 



144 X 0.0361 



= 4.45 inches. 



(0.074495 = density of air at 75 degrees F. and 0.0361 = pressure pro- 
duced by one inch of water in pounds per square inch.) 

If the discharge orifice be opened to its maximum (Fig. 174) the 
static pressure indicated by manometer B becomes zero, since there is 



s^\ 



& 



(A) 
Orifice Wide Open 



(B) 



Fig. 174. 

no resistance due to the air flow, while the water in A stands at a height 
H the exact equivalent of the velocity head in accordance with the 
hydraulic formula, 

v = VYtfT, (83) 

in which v is the velocity of the air in feet per second. 

If the orifice be partially closed, say 50 per cent, as in Fig. 175, B indi- 
cates the static pressure, while A gives the dynamic or total pressure 



^\ 



c> 



rp\ 



^s£ 



(A) 



(B) 



Orifice partly closed 



^\ 






(C) 



6> 



Fig. 175. 



due to both velocity and resistance. Tne difference between A and B 
is therefore the pressure due to velocity alone. By connecting the 



294 STEAM POWER PLANT ENGINEERING 

two manometers as indicated in Fig. 175C the velocity pressure is 
given directly.* 

Pressure. — According to Murgue's theory the maximum pressure 
which may be developed by a blower or exhauster varies with the 
square of the speed and may be expressed 

P=— -> (84) 

y 
in which 

p = pressure, pounds per square foot, 
8 = density of the air, pounds per cubic foot, 
u = peripheral velocity, feet per second, 
C = a coefficient obtained by experiment. 

Tables 53 and 54 give the relationship between pressure and speed 
for various sizes of forced and induced-draft fans. 

Fig. 178 shows the relationship between pressure and speed in a 
45-inch Buffalo blower as tested at the Armour Institute of Technology. 

Velocity of Discharge. — The maximum velocity of the air leav- 
ing the tips of the blades varies directly as the peripheral speed, 

V = Ku, (85) 

in which 

V = velocity of the air discharged, feet per second, 
K = a coefficient obtained by experiment, 
u = peripheral velocity, feet per second. 

For practical purposes the velocity of discharge with outlet wide 
open may be assumed to be that of the periphery. 

Capacity. — The relationship between capacity and speed, capacity 
and discharge opening for a 45-inch pressure blower is given in Figs. 177 
and 178. 

As will be noted, the capacity varies almost directly with the speed of 
the wheel and the area of discharge as expressed by the equation 

Q = BttADN, (86) 

in which 

Q = cubic feet discharge per minute, 

B = coefficient determined from experiment, 

A = area discharge opening, square feet, 

D = diameter of the wheel, 

N = r.p.m. of the wheel. 

* The manometer readings in C, Fig. 175, indicate the velocity pressure for 
the point D. For a method of determining the average velocity of the conduit at a 
section through D, see Eng. News, Dec. 21, 1905, p. 660. 



MECHANICAL DRAFT 295 

Power. — The power required to drive a fan is proportional to the 
cube of the speed, 

Horse power = XAN\ (87) 

in which 

X = a coefficient determined by experiment, 

A = area discharge outlet, square feet, 

N = r.p.m. 

The marked increase in power required for even a moderate increase 
in speed should be borne in mind in selecting a fan. (See power curves, 
Fig. 178.) It is, as a rule, more economical to err in selecting too large 
a fan than one which must be forced above its rated capacity. 

Manometric Efficiency. — This efficiency is the ratio of the dy- 
namic head as actually observed to the maximum theoretical dynamic 
head, or 

^man = 77' (88) 

in which h is determined from the actual manometer reading and H is 
calculated from equation (82). 

Volumetric Efficiency. — This is the ratio between the actual 
volume of air passing in a given time divided by the impeller dis- 
placement for the same period, or 

Evo1== ^dWb } (89) 

in which 

Q = volume discharged, cubic feet per minute, 
D = diameter of the impeller, feet, 
B = width of the impeller, feet, 
N = r.p.m. 

Mechanical Efficiency, or simply fan efficiency is the ratio of the 
total work done by the fan in moving the air to the horse-power input 
to the fan, or 

_ Q'h 

^ mec " HiX 33,000' m) 

in which 

Q' = weight discharged, pounds per minute, 
h = dynamic head, feet of air, 
Hi = horse-power input. 

In practice the size of fan is proportioned upon experience rather 
than theory, the usual procedure necessitating the use of curves based 
upon the performance of fans of the type under consideration. 



296 



STEAM POWER PLANT ENGINEERING 



The curves in Fig. 176 were computed by Mr. F. R. Still of the 
American Blower Company, and give the performance of steel-plate 
fans as manufactured by this company. These curves apply to this 
type and make of fan only, though the difference is not very great 
for any type of centrifugal fan. The "ratio of opening" refers to the 
actual percentage of opening compared with the total discharge. The 



















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10 20 30 40 50 60 70 80 90 100 110 120 130 140 
Ratio of Effect Per Cent 
Fig. 176. Performance of Steel Plate Fans. 

" ratio of effect" is the relative effect produced by restricting the dis- 
charge. The abbreviations are as follows: 
D.P. = dynamic or total pressure. 
P.V.P. = pressure created by a column of air moving at the same 
velocity as the periphery. 
S.P. = static pressure. 
V.P. = velocity pressure = D.P. — S.P. 

Suppose a fan with an unrestricted inlet and outlet delivers 25,000 
cubic feet of air per minute against a head (D.P.) of 0.33 inch with a 
peripheral velocity requiring 6.16 horse power. It appears from the 
curves that if the discharge outlet is restricted to 50 per cent of the 
full area, only 12,500 cubic feet will be delivered; the pressure will be 
increased to 1.03 inches, and the power required drops to 4.84 horse 
power. If the outlet be still further reduced to 20 per cent of the full 
opening the capacity will drop to 5000 cubic feet, the pressure will 
increase to 1.15 inches, and the power will be decreased to 3.45 horse 



MECHANICAL DRAFT 297 

power. With a discharge area of 60 per cent, the mechanical efficiency 
is a maximum, and equal to about 43 per cent. With orifice closed 
the horse power required to drive the fan is about 37 per cent of that 
required when discharging the maximum volume of air. 

Curve "K" in Fig. 176 was determined from the empirical formula 
(based upon Murgue's theorem) 

A-&, (90a) 

in which 

A = area of the inlet orifice, square feet, 

Q = volume of gas, thousands of cubic feet per minute, 

P = draft at the inlet in inches of water, 

K = constant determined by experiment. 

The curves in Figs. 177 and 178 are plotted from tests made at the 
Armour Institute of Technology on a 45-inch Buffalo pressure blower, 
and are characteristic of this type of fan. 

Measurement of Air in Fan Work: C. H. Treat, Jour. A.S.M.E., Sept., 1912, 
p. 1341. Some Experiences with the Pitot Tube on High and Low Air Velocities: 
F. H. Kneeland, Jour. A.S.M.E., Nov., 1911, p. 1407. Experiments with Ventilat- 
ing Fans and Pipes: Capt. D. W. Taylor, Soc. Naval Arch, and Marine Engrs., 
1905, p. 35. The Measurements of Gases: Carl C. Thomas, Jour. Frank. Inst., 
Nov., 1911, p. 411. Experiments with the Pitot Tube in Measuring the Velocities of 
Gases: R. Burnham, Eng. News, Dec. 21, 1905, p. 660. Pressure Fans vs. Exhaust 
Fan: Bulletin Am. Inst. Min. Engrs., Feb., 1909. 

168. Determination of Size of Fan. — The following analysis, based 
upon a paper on Mechanical Draft by F. R. Still of the American 
Blower Company, gives a good idea of the usual procedure in deter- 
mining the size of fan for an induced draft installation. (Jour. West. 
Soc. Engr., May, 1902.) 

Example: Determine the size of induced fan and the approximate 
power required to drive it, for a boiler plant rated at 1000 horse power; 
temperature of flue gases 500 degrees F.; heat value of coal 14,000 
B.t.u. per pound; ash 5 per cent; draft required, 1 inch of water 
pressure. 

Assuming a boiler efficiency of 70 per cent, the evaporation will be 

Q ' X 0.70 =10.15 pounds of water from and at 212 degrees F. 

per pound of coal. 

Since one boiler horse power is equivalent to the evaporation of 
34.5 pounds of water per hour from and at 212 degrees F., the evapora- 
tion per hour will be 34.5 X 1000 = 34,500 pounds, and the coal burned 

per hour, 34 500 _ nn 

., ' ir = 3400 pounds. 
10.15 



298 



STEAM POWER PLANT ENGINEERING 



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Area.of Discharge Opening, Square Feet 
Fig. 177. 







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Buffalo Blower 

Discharge Area Constant 

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300 900 1000 ,1100 
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Fig. 178. 



1200 1300 1400 * 1500 



MECHANICAL DRAFT 299 

Allowing 18 pounds of flue gas per pound of combustible, 5 per cent 
for ash and 5 per cent for leaks, the fan will have to handle, at 500 
degrees F., approximately 20 X 3400 = 68,000 pounds of gas per hour, 
or 26,000 cubic feet per minute. It is customary, when little is known 
about a plant in which a fan is to be installed, to assume that the re- 
sistance is equivalent to restricting the discharge outlet 25 per cent. 
Hence, in this problem the various factors are referred to a " ratio of 
opening" of 75 per cent (see Fig. 176). 

From formula 90a, the area of the inlet should be 

A = —7= = — — = 12.6 square feet, 

which corresponds to a diameter of 48 inches. (K = 0.485 is taken 
from the curves in Fig. 176.) 

The area of the inlet should not exceed 40 per cent of the area of the 
side of the wheel; the latter, then, will be 

-rr-j- = 31.5 square feet, 

which corresponds to a diameter of 76 inches (6.3 feet). 

Referring to Fig. 176, the ratio of dynamic pressure to peripheral 
velocity pressure (D.P. to P.V.P. at 75 per cent opening) is 0.73. The 

peripheral velocity pressure will be ^-=5 = 1.37 inches of water. 

The peripheral velocity is 

U = V2 gH' = 8.03 y/H', where H' is the peripheral velocity pres- 
sure expressed in feet of gas, or, 

since H' = . Q ^ ' , where p = inches water, 

U = 87.5 VL37 

= 102.5 feet per second 
= 6150 feet per minute. 

The maximum effective discharge area, which an inclosed fan of this 
type may have, and still maintain the pressure equivalent of the pe- 
ripheral velocity, is usually called the " blast area." With a larger area 
the pressure will be reduced, but with a smaller area will remain sub- 
stantially constant. The velocity of the discharge is practically that 

of the tips of the blades, whence the blast area is equal to n ' rri = 4.23 

0150 

square feet, which, with this type of fan, is found to be about J the 

projected rectangle of the wheel, therefore, 

The projected rectangle = 4.23 X 3 = 12.7 square feet. 



300 STEAM POWER PLANT ENGINEERING 

The proper width of periphery is found by dividing this area by the 

wheel diameter, thus, 

12 7 
width of blades = -^ = 2.02 feet = 24.2 inches, 
6.3 

and speed of fan = Q =311 r.p.m. 

o.l4 X O.o 



w 



Volume of gas (cu. ft. per min.) X Pressure (lb. per sq. ft.) 
33,000 X efficiency of fan 
Tjr 26,000X5.2 1AOU . , 
W = 33,000 X 0.4 = 10 - 2 bmke h0rse P ° Wer - 



(5.2 = pressure in pounds per square foot equivalent to one inch of 
water, and 0.4 is the mechanical efficiency for 75 per cent opening as 
taken from curve in Fig. 176.) 

Assuming a steam consumption of 70 pounds per brake horse power 
for a small, simple non-condensing high-speed engine, the steam con- 
sumed per hour will be 

10.2 X 70 = 714 pounds per hour, or 2.3 per cent of the total steam 
capacity of the boilers. 

Table 53 gives the capacity and horse power required for various 
sizes of forced-draft fans, and Table 54 gives similar data for induced- 
draft fans. 

169. Chimney vs. Mechanical Draft. — The choice of chimney or 
mechanical draft depends largely upon local conditions. Many power 
plants with tall stacks are provided with forced-draft apparatus to be 
used in emergencies, but, as a general rule, where ordinances require 
high chimneys mechanical draft is not considered. In a few isolated 
cases stokers of the forced-draft type are used in connection with 
chimneys as high as 250 feet, but such installations are limited to large 
central stations with heavy peak loads. 

Where there are no limitations to the height of stack, mechanical 
draft offers many advantages over chimney draft, especially for rail- 
road work and large lighting plants. With certain types of grates and 
for low-grade fuels and anthracite culm or dust, it is indispensable. 
Again, where a fair quality of fuel is obtainable the size of plant may 
determine the choice. 

First Cost: In small plants of, say, 100 to 150 horse power the cost 
of a guyed steel chimney, 75 feet in height or less, would be consider- 
ably less than that of a mechanical-draft system, and once erected would 
cost practically nothing for operation, while the power required to oper- 
ate a fan in so small a plant would amount to 5 per cent or more of the 
total steaming capacity. 



MECHANICAL DRAFT 



301 



TABLE 53. 

CAPACITIES OF FORCED-DRAFT FANS. 
(Steel Plate Fans.) 



For Forced Draft, Temperature of Air 60°. 











Pressure 


in Inches of Water. 






Cubic Feet 
of Air De- 






























Diam- 




0.5 


0.75 


1.00 


1.25 


1.50 


2.00 


2.50 


eter of 


Furnace 
















Fan. 






























per 


§ 


Pn 


3 


Ah' 


ez 


Ph 


§ 


Ph" 


f^ 


Ph" 


f^ 


Ph 


S 


Ph* 




Minute. 


Ph 

pi 
510 


1.6 


Ph 
560 


1.8 


Ph 
600 


1.9 


Ph 
Ph 

640 


Ph' 

2.1 


Ph 

710 


2.3 


Ph 

pi 
780 


m 

2.5 


Ph 

p4 

850 


w 


2' 6* 


4,200 


2.7 


3' 


5,800 


430 


2.2 


460 


2.4 


490 


2.6 


530 


2.8 


590 


3.1 


640 


3.4 


710 


3.8 


3' 6" 


7,800. 


360 


3.0 


400 


3.3 


420 


3.5 


450 


3.8 


500 


4.2 


550 


4.6 


610 


5.1 


4' 


10,000 


320 


3.9 


350 


4.2 


370 


4.4 


400 


4.9 


440 


5.4 


480 


5.9 


530 


6.5 


4' 6" 


12,400 


290 


4.8 


310 


5.2 


330 


5.6 


360 


6.0 


400 


6.7 


430 


7.3 


470 


8.0 


5' 


15,200 


250 


5.9 


270 


6.4 


290 


6.8 


310 


7.4 


350 


8.2 


380 


8.9 


420 


9.8 


5' 6* 


18,200 


230 


7.0 


250 


7.7 


270 


8.2 


300 


8.8 


330 


9.8 


360 


10.6 


390 


11.8 


6' 


21,400 


210 


8.3 


230 


9.1 


250 


9.6 


260 


10.4 


290 


11.5 


320 


12.5 


350 


13.9 


7' 


28,800 


180 


11.2 


200 


12.2 


210 


13.0 


230 


14.0 


250 


15.5 


280 


16.8 


300 


18.7 


8' 


37,200 


160 


14.4 


170 


15.7 


190 


16.7 


200 


18.1 


220 


20.1 


240 


21.8 


270 


22.5 


9' 


46,800 


140 


18.1 


160 


19.8 


170 


21.1 


180 


22.7 


200 


25.3 


220 


27.4 


240 


30.3 


10' 


57,400 


130 


22.2 


140 


24.3 


150 


25.8 


160 


27.9 


180 


3.1 


200 


33.6 


210 


37.2 



Discharge velocity 2000 feet per minute. 
TABLE 54. 

CAPACITIES OF INDUCED-DRAFT FANS. 
(Steel Plate Fans.) 









For Induced Draft, Temp. 


of Flue Gases 500°. 












Cubic Feet 






Pressure 


in Inches of Water. 




of Air at 
60°Temp. 












Diam- 


0.5 


0.75 


1.00 


1.25 


1.50 


2.00 


2.50 


eter of 


Drawn 
into Fur- 
















Fan. 
































nace per 


S 


Ph' 


3* 


Ph 


a 


Ph' 


% 


(U 


B 


Ph' 


a 


(^ 


s 


Ph* 




Minute." 


Ph' 
Ph 

688 


W 

2.2 


Ph' 
Ph 

756 


W 

2.4 


Ph' 

pi 
810 


W 

2.6 


Ph 
Ph 

864 


m 

2.8 


Ph' 

pi 
958 


3.1 


Ph' 

pi 
1053 


m 

3.4 


Ph' 

p4 

1147 


w 


2' 6" 


3,000 


3.6 


3' 


4,200 


580 


3.0 


621 


3.2 


661 


3.5 


715 


3.8 


796 


4.2 


864 


4.6 


958 


5.1 


3' 6" 


5,700 


486 


4.0 


540 


4.5 


567 


4.7 


607 


5.1 


675 


5.7 


742 


6.2 


823 


6.9 


4' 


7,300 


432 


5.3 


472 


5.7 


500 


6.1 


540 


6.6 


594 


7.3 


648 


8.0 


715 


8.8 


4' 6" 


9,300 


390 


6.5 


418 


7.0 


445 


7.5 


486 


8.1 


540 


9.0 


580 


9.8 


634 


10.8 


5' 


11,100 


337 


8.0 


364 


8.6 


391 


9.2 


418 


10.0 


472 


11.1 


513 


12.0 


567 


13.2 


5' 6* 


13,300 


310 


9.5 


337 


10.4 


364 


11.1 


405 


11.9 


445 


13.2 


486 


14.3 


526 


15.9 


6' 


15,600 


283 


11.2 


310 


12.3 


337 


13.0 


351 


14.0 


391 


15.5 


432 


16.9 


472 


18.7 


V 


21,000 


243 


15.1 


270 


16.5 


283 


17.5 


310 


18.9 


337 


20.9 


378 


22.6 


405 


25.2 


8' 


27,100 


216 


19.4 


230 


21.2 


256 


22.5 


270 


24.4 


297 


27.1 


324 


29.4 


364 


30.4 


9' 


34,200 


189 


24.4 


216 


26.7 


230 


28.5 


243 


30.6 


270 


34.1 


297 


37.0 


324 


40.9 


10' 


41,900 


175 


30.0 


190 


32.8 


202 


34.8 


216 


37.6 


243 


41.8 


270 


45.3 


283 


50.2 



302 



STEAM POWER PLANT ENGINEERING 



TABLE 55. 

CAPACITIES OF FORCED-DRAFT FANS.* 
(Sirocco Type.) 

(Figures given Represent Dynamic Pressures in Ozs. per Sq. In. For Static Pressure Deduct 28.8 Per Cent. 
For Velocity Pressure Deduct 71.2 Per Cent.) 







1 Oz. 


*Oz. 


lOz. 


lOz. 


UOz. 


HOz. 


If Oz. 


2 0z. 


2£Oz. 


3 0z. 


6 


Cu. Ft. 
R.P.M. 
B.H.P. 


155 

1,145 

0.0185 


220 
1,615 
0.052 


270 
1,980 
0.095 


310 
2,290 
0.147 


350 
2,560 
0.205 


380 
2.800 
0.270 


410 
3,025 
0.34 


440 
3,230 
0.42 


490 
3,616 
0.58 


540 
3,960 
0.76 


12 


Cu. Ft. 
R.P.M. 
B.H.P. 


625 

572 

0.074 


880 

808 

0.208 


1,080 

990 

0.381 


1,250 
1,145 

0.588 


1,400 
1,280 
0.82 


1,530 
1,400 
1.08 


1,650 
1,512 
1.36 


1,770 
1,615 
1.66 


1,970 

1,808 
2.32 


2,170 
1,980 
3.05 


18 


Cu. Ft. 
R.P.M. 
B.H.P. 


1,410 

381 

0.167 


1,990 

538 

0.470 


2,440 

660 

0.862 


2,820 
762 
1.33 


3,160 
850 

1.85 


3,450 

933 

2.43 


3,720 
1,010 
3.07 


3,980 
1,076 
3.75 


4,450 
1,204 
5.25 


4,880 

1,320 

6.9 


24 


Cu. Ft. 
R.P.M. 
B.H.P. 


2,500 

286 

0.296 


3,540 

404 

0.832 


4,340 
495 
1.53 


5,000 

572 

2.35 


5,600 

640 

3.28 


6,120 

700 

4.32 


6,620 

756 

5.44 


7,080 

807 

6.64 


7,900 
904 
9.3 


8,680 
990 
12.2 


30 


Cu. Ft. 
R.P.M. 
B.H.P. 


3,910 

228 

0.460 


5,520 
322 
1.30 


6,770 

395 

2.40 


7,820 

456 

3.68 


8,750 

510 

5.15 


9,600 

560 

6.75 


10,350 

604 

8.53 


11,050 
645 
10.4 


12,350 
722 
14.5 


13,550 
790 
19.1 


36 


Cu. Ft. 
R.P.M. 
B.H.P. 


5,650 

190 

0.665 


7,950 
269 

1.87 


9,750 

330 

3.44 


11,300 

381 

5.30 


12,640 
425 
7.40 


13,800 

466 

9.72 


14,900 

504 

12.25 


15,900 
538 
15.0 


17,800 

602 

20.9 


19,500 

660 

27.5 


48 


Cu. Ft. 
R.P.M. 
B.H.P. 


10,000 
143 

1.18 


14,150 

202 

3.32 


17,350 

248 

6.10 


20,000 

286 

9.40 


22,400 
320 
13.1 


24,500 
350 
17.2 


26,500 

378 

21.75 


28,300 

403 

26.6 


31,600 
452 
37.1 


34,700 

495 

48.8 


60 


Cu. Ft. 
R.P.M. 
B.H.P. 


15,650 
114 

1.84 


22,100 

161 

5.20 


27,100 

198 

9.58 


31,300 
228 
14.7 


35,000 

255 

20.6 


38,400 

280 

27.0 


41,400 

302 

34.1 


44,200 

322 

41.6 


49,400 

361 

58.2 


54,200 

396 

76.5 


72 


Cu. Ft. 
R.P.M. 
B.H.P. 


22,600 

95 

2.66 


31,800 
134 

7.48 


39,000 

165 

13.7 


45,200 

190 

21.2 


50,600 

212 

29.6 


55,200 

233 

3S.9 


59,600 

252 

49.0 


63,600 

269 

59.8 


71,200 

301 

83.6 


78,000 
330 
110 


84 


Cu. Ft. 
R.P.M. 
B.H.P. 


30,800 

81 

3.61 


43,400 

115 

10.2 


53,200 
142 

18.7 


61,600 

163 

28.9 


68,700 

182 

40.4 


75,200 

200 

53.0 


81,200 

216 

66.8 


86,800 

231 

81.7 


97,100 
258 
114 


106,400 
283 
150 


90 


Cu. Ft. 
R.P.M. 
B.H.P. 


35,250 

76 

4.14 


49,800 

107 

11.7 


61,000 

132 

21.5 


70,500 

152 

33.1 


78,800 

170 

46.2 


86,400 

186 

60.7 


93,300 
201 
76.7 


99,600 

214 

93.6 


111,200 
241 
131 


122,000 
264 
172 



* A number of sizes have been omitted. 



A tall, self-supporting chimney for larger plants, however, is very 
costly as compared with a fan system of equal capacity. For example, 
a brick chimney 175 feet high and 10 feet in diameter, foundation and 
all, capable of furnishing the necessary draft for a 3000-horse-power 
plant, will cost about $10,000. A two-fan induced system of equiva- 
lent capacity will cost in the neighborhood of $5000, a one-fan system 
$3500, and a forced-draft system $2500. See Fig. 179. With interest 
at 5 per cent, depreciation 5 per cent, taxes 1 per cent, and insurance 
one-half /per cent, the annual fixed charges will be $575, $402.50, $287.50 
respectively, for the fan equipment. 



MECHANICAL DRAFT 



303 



Depreciation and Maintenance: The depreciation of a well-designed 
masonry or concrete stack is very low, and 2 per cent is a liberal factor. 
Maintenance is practically negligible, as it requires no attention what- 
ever for years. A steel stack, however, must be kept well painted or 
corrosion will take place rapidly. The depreciation and maintenance 
charges on a mechanical-draft system will range from 4 per cent to 10 
per cent of the original outlay. 



15 


















































I 










































































14 


























































































































































































12 












































































































































































I 










































































10 


























































































































































/ 


^ 




























«. D 
































/ 






























o 

O 8 

o 






























/ 




























































/ 
































+? ., 




























/ 


































9 


























V 


1 


































tf 
























$ 


p 


































6 






















<f/ 




































5 






















/ 




























































/ 


























^ 














4 


















/ 


/ 






































































22 





a ] 


fax 


£- 












/ 












3 


























tS 


^ 












A 1 


fe? 


v,- 


' 




















A 


s 












s<& 


X* 














^ 


»S 


1-- 






1 




















/ 




















y 


•^ 




i^ 


*& 














I 


















/ 


/ 


; 
















/\ 




'"- 


-^ 




._- 




•e,d 


D 


ra 


r tT 


— 




































y 














"E 


























1 






- 


— 


— ' 






__. 


— - 


- 


' 
















































""" 























































1000 



2000 



3000 



Horse Power 
Fig. 179. Comparative Costs of Chimneys and Mechanical Draft. (W. B. Snow.) 



Cost of Operation: Once erected, the comparative cost of operating 
a chimney is practically nothing; that is, of course, on the assumption 
that the chimney and fan exhaust equal volumes of gas per pound of 
fuel and at the same temperature. A fan system requires for its opera- 
tion from one and one-half per cent to five per cent of the total steam- 
ing capacity of the plant, depending upon the type and character of 
the fan engine or motor, and the conditions of operation. 



304 



STEAM POWER PLANT ENGINEERING 



Efficiency: With fan draft a very thick fire can be maintained on 
the grate, thus permitting a high rate of combustion, and minimum 
air per pound of fuel, both of which result in increased boiler efficiency. 
The influence of the rate of combustion on air supply in a specific case 
is illustrated in Fig. 180. For the same temperature of discharge each 
pound of air in excess of theoretical requirements results in a loss of 
about one per cent of the total heat in the fuel. With fan draft an 
average figure is 18 pounds of air per pound of bituminous coal against 
24 pounds for the chimney, a saving of 5 per cent in favor of the fan. 
Again, a fan permits of a low temperature of the flue gases without 
affecting the draft, while lowering the temperature in the chimney 



gOO 



200 



^100 



O 



10 



20 



-10 



50 



Fig. 180. 



Lb.Coal Burned Per Sq.Ft.Grate Per Hr. 
Influence of Rate of Combustion on Air Supply. — Forced Draft. 



reduces the draft as shown in Table 42. From Table 11 we see that 
a reduction in flue gas temperature of 25 degrees F. will increase the 
boiler efficiency about one per cent. With an economizer the flue 
gases may be reduced to 350 degrees F., with a net saving of about 
500 — 350 = 150, or 6 per cent of the total fuel. It is in this con- 
nection that the fan draft is peculiarly suitable. Of course, the chimney 
may be provided with an economizer, effecting the same reduction in 
temperature, but its height must be made sufficiently great to overcome 
the additional resistance of the economizer and the reduction in tem- 
perature of the chimney gases. 

Flexibility: With a fan the draft may be readily regulated for 
sudden increased or decreased requirements, independent of the boiler 
performance. Damp and muggy days appreciably affect the draft of 
a chimney, as do adverse air currents and high winds. 



MECHANICAL DRAFT 



305 



mmm. 



jfv- 



IUU3JS »>BIV WIOJJ 8diJ 

janojqwoo jaduioa 



oj ad; j aanssaa^ 




t^ -'" / ^■-J3p.^;5^::- - 



306 STEAM POWER PLANT ENGINEERING 

Smoke: Smokeless combustion is more readily effected with arti- 
ficial draft than with natural draft, as a thicker fire can be carried, and 
the correct proportion of air can be more readily adjusted. 

170. Balanced Draft. — Fig. 181 illustrates an application of the 
McLean " Balanced Draft" system to a water-tube boiler. The 
equipment consists of a blower, the speed of which is regulated by the 
steam pressure, so that the draft in the fire box is maintained at ap- 
proximately atmospheric pressure. The chief claims for this system 
are: (1) the velocity of the gases over the tubes is reduced, and short 
circuiting is prevented; (2) the correct proportion of air to fuel is readily 
maintained; (3) infiltration of air through the setting is impossible, 
as the pressures are " balanced"; (4) sudden changes in load are cor- 
rectly taken care of. Tests of the apparatus at the Fuller Building, 
New York, gave excellent results (Trans. A.S.M.E., 26-641). 



CHAPTER IX. 

RECIPROCATING STEAM ENGINES. 

171. Introductory. — The type of prime mover best suited [for a 
given installation is the one which delivers the required power at the 
lowest cost, taking into consideration all charges, fixed and operating. 
These include not only the cost of fuel, labor, supplies and repairs, 
but all overhead charges such as interest on the investment, deprecia- 
tion, maintenance and taxes. Space requirements and continuity of 
operation are often of vital importance, and may greatly influence 
the selection of type of prime mover and auxiliary apparatus. In 
many situations the gas engine and producer are productive of the 
highest commercial economy; in others the choice lies between the 
reciprocating steam engine or turbine, occasionally the hydroelectric 
plant offers the best returns, but in general each proposed installation 
is a problem in itself, and general rules are without purpose. 

The reciprocating steam engine is the most widely distributed prime 
mover in the power world, and although its field of usefulness has been 
greatly encroached upon in recent years by the steam turbine and gas 
engine it is still an important heat engine and will probably continue 
to be a factor for years to come. In a general sense the piston engine 
is superior to the turbine for variable speed, slow rotative speeds and 
heavy starting torque, while the turbine has practically superseded the 
engine for large central station units and for auxiliaries requiring high 
rotative speed. From a purely thermal standpoint the internal combus- 
tion engine is vastly superior to the steam engine and the turbine is 
more economical in space requirements, but taking into consideration 
all of the items affecting the production of power, the reciprocating 
engine may still prove to be the better investment in many situations. 

172. The Ideal Engine. — In every heat engine the working fluid 
goes through a circuit or cycle of operation. Beginning at a particular 
condition it passes through a series of successive states of pressure, 
volume and temperature and returns to the initial condition. An 
ideally perfect engine which effects the highest possible conversion of 
heat into mechanical work for a given cycle is taken as a standard of 
comparison for the performance of the actual engine. Two such 
standards are adopted in connection with the steam engine, (1) the 

307 



308 



STEAM POWER PLANT ENGINEERING 



ideal engine operating in the Carnot cycle, and (2) the ideal engine 
operating in the Rankine cycle. 

173. The Carnot Cycle. — The diagrams in Fig. 182 represent the 
action of the working fluid in an ideal steam engine cylinder, operating 
in the Carnot cycle. (A) represents the familiar indicator card or 
pressure-volume diagram, and (B) the temperature-entropy diagram. 
The former illustrates the kinetic action of the steam in the cylinder 
and the latter the thermal action. At the beginning of the stroke, a, 
the non-conducting cylinder contains a mixture of steam and water at 
temperature T± and pressure pi. Heat is applied at temperature T\ 
and pressure pi until a part or all of the liquid is vaporized and the 
isothermal expansion forces the frictionless piston to position b. From 




T 


I 






\ 






a 


A 




wl 


Mi 


v 




7 


i 


■i 


\ 




r 


T, t 


c \ 


a 










H 




' 


(B) l - 





Volume 



«i b. 

Entropy 



Fig. 182. The Carnot Cycle (Saturated Steam). 

b to c the vapor expands adiabatically and forces the piston to the 
end of its stroke. On the return stroke from c to d the working fluid 
is compressed isothermally with condensation and rapid rejection of 
heat. The cycle is completed by an adiabatic compression from d 
to a. * 

If x = the quality of the steam at the point indicated by the sub- 
script, 
r = latent heat of vaporization at the pressure indicated by the 

subscript, B.t.u. per pound of steam, 
T = absolute temperature of the steam, degrees F. 

The heat Hi absorbed by one pound of the mixture in passing from 

a to b is 

H x = nx b — nx a . (91) 

The heat H 2 rejected in passing from c to d is 

H 2 = r 2 x c — r 2 x d . (92) 

* It is commonly assumed that there is only water in the cylinder at a and satu- 
rated steam at b, but there is no necessity for such an assumption, and it in no way 
affects the efficiency. 



STEAM ENGINES 309 

The heat H c converted into work during the cycle is 
H c = Hi — H2 

= ri{x b — x a ) — r 2 (x c — x d ). (93) 

From thermodynamics we have as the adiabatic equation for a liquid 

and its vapor 

T 

r 2 (x c — x d ) = Yn(x b — Xa). (94) 

Substituting this value of r 2 (x c — x d ) in equation (93) we get 

T — T 
H c = r x {x h - x a ) — ^ — -• (95) 

The efficiency E c of the perfect engine operating in the Carnot cycle 

is 

( _ Ji"^ 
= H,-H 2 = n{Xb Xa) T x = T x - T 2 (96) 

H x nx b — nx a T 1 

which is independent of the nature of the working substance and de- 
pendent only on the range in temperature. The upper limit of tem- 
perature is that corresponding to boiler pressure, and the lower limit 
to that of the exhaust steam. Evidently the greater this temperature 
range the more nearly does this efficiency approach unity, but with the 
present limits of temperature used in steam engines it cannot exceed 
38 per cent. 

In the wholly ideal Carnot cycle the entire cycle — heat reception 
and expansion, heat rejection and compression — is supposed to be 
performed within the cylinder itself, using an unchanged body of work- 
ing substance over and over again. While not absolutely impossible 
this manner of operation is commercially impracticable. ^ 

The nearest approach of any actual engine to the Carnot cycle is 
accomplished by the Nordberg system of progressive feed-water heat- 
ing, in which the water is successively heated from the receivers inter- 
mediate between each pair of cylinders. (For a description of this 
engine see Eng. News, May 4, 1899, p. 283.) 

Example: Determine the efficiency of the ideal engine working in 
the Carnot cycle if the boiler pressure is 200 pounds absolute and the 
back pressure 2 pounds absolute. 

T x = 388 + 459.6 = 847.6. 
T 2 = 126.1 + 459.6 = 585.7. 
847.6 - 585.7 



E = 



847.6 
0.309 or 30.9 per cent. 



310 



STEAM POWER PLANT ENGINEERING 



174. The Rankine Cycle with Complete Expansion. * — The Carnot 
cycle is practically impossible for an engine using superheated steam 
at constant pressure, and in general is not closely simulated by an 
engine using saturated steam. It represents, however, the theoretical 
limit of perfection of any heat engine. 

The diagrams in Fig. 183 represent the action of the working fluid in 
an ideal engine cylinder, operating in the Rankine cycle, which closely 
parallels the cycle of the actual engine, ab represents the admission 
of steam from the boiler at pressure pi) be is an adiabatic expansion 
to exhaust pressure p 2 ; cd represents the exhaust, and da is an adiabatic 
compression to the initial pressure. 




T 


xvPa 

J! 


!<* 




b 




2 

3 




HP 


\ 


1 
o 

a 

o 
H 


/ 


' 


i 

(B) ^ 


\ 



Volume 



Entropy 



N 



Fig. 183. The Rankine Cycle with Complete Expansion (Saturated Steam). 

Let Hi = the total heat of one pound of steam at pressure p lt 

H2 = total heat of one pound of steam at pressure p 2 after adia- 
batic expansion from pressure pi. 
q 2 = heat of the liquid at pressure p 2 . 



The heat changed into work per pound of steam is 

Hi — H 2 . 



(97) 



The heat necessary to raise the feed water from the temperature of 
exhaust to the temperature in the boiler and evaporate it is 

Hi - q 2 .' (98) 

The efficiency, E r , of the cycle, or the ratio of the heat equivalent of 
the useful work to the heat supplied, is 

Hi — H 2 



E r = 



Hi — 32 



(99) 



* This is often called the Clausius cycle since it was published simultaneously, 
but independently, by both Clausius and Rankine. This cycle has been adopted 
by the British Society of Civil Engineers, and is generally accepted in this country 
as the standard of comparison for steam engines and turbines. 



STEAM ENGINES 



311 



»«D 




312 



STEAM POWER PLANT ENGINEERING 



and the water rate of steam consumption of the perfect engine, W, 
pounds per horse-power hour, may be expressed as 

2546 



W = 

For dry steam Hi = n-\- q h 

For wet steam Hi = Xi r h + qi, 

For superheated steam Hi = n + qi + Cit 8} 




(100) 

(101) 
(102) 
(103) 



Fig. 185. Typical Piston Engine, Single Cylinder, Automatic Governor. 

in which 

r*i = heat of vaporization at pressure pi, 
Xi = quality of steam at pressure p h 
qi = heat of the liquid at pressure p h 

Ci = mean specific heat of the superheated steam at pressure p h 
t 8 = degree of superheat or the difference in temperature between the 
superheated and saturated steam at pressure pi. 



STEAM ENGINES 



313 



If the steam after adiabatic expansion is wet as is the usual case 

H 2 = x 2 r 2 + q 2 . (104) 

If initial superheat is so high that the steam at the end of expansion 

is still superheated 

H 2 = r 2 + q 2 + CV, (105) 

in which 

x 2 , 7*2, q 2 — quality, heat of vaporization and heat of the liquid re- 
spectively, at pressure p 2 , 
C = mean specific heat of superheated steam at pressure p 2 , 
t s f = degree of superheat at pressure p 2 . 

Before the total heat-entropy or Mollier diagram came into common 
use it was necessary to calculate H 2 , x 2 , and C from thermodynamic 
equations, a tedious and laborious procedure 
and particularly so with highly superheated 
steam. With the aid of this diagram all prob- 
lems involving adiabatic expansion are solved 
with ease and accuracy; in fact, the Mollier 
diagram has to all intents and purposes sup- 
planted the steam tables in this connection. 
For this reason, the thermodynamic equations 
for solving H 2 , x 2 , and C will be omitted. See 
appendix L. 

Example: Determine the efficiency and water 
rate of the ideal engine working in the Rankine 
cycle if the steam pressure is 200 pounds abso- 
lute, superheat 250 degrees F., and exhaust pressure 0.5 pounds absolute 

From steam tables or the Mollier diagram we find 

H 1 = 1332 B.t.u. and q 2 = 48 B.t.u., 

From the Mollier diagram we find 

H 2 = 908 B.t.u., 

„ 1332 - 908 _ 00 

E r = 7777^ j^r = 0.33 or 33 per cent, 




Entropy 

Fig. 186. The Rankine Cycle 

for Superheated Steam. 



W r 



1332 - 48 

2546 
1332 - 908 



6.0 pounds per horse-power hour. 



175. The Rankine Cycle with Incomplete Expansion. — If expansion 
after cut-off is not carried far enough to reduce the pressure to that of 
the back pressure line, as shown in Fig. 187, the Rankine cycle more 
nearly simulates the cycle of the actual engine. This cutting off the 
" toe " of the diagram decreases the ideal efficiency, but permits of the 
use of a smaller cylinder. 



314 



STEAM POWER PLANT ENGINEERING 



The work during admission is 

Wi = pi (xiiii + a) foot-pounds, (106) 

in which 

Ui = increase of volume due to vaporization of a pound of 

steam, 
o- = specific volume of water. 

The work during expansion from b to c is 

W 2 = (xipi + qi — x c p c — q c ) X 778 foot-pounds, (107) 

in which 

x, p, and q are the quality, heat equivalent of the internal work 
during vaporization and the heat of the liquid respectively at pressures 
indicated by the subscript. 

t _/ 



p 


a 


b 


t 

s 

a: 
XD 

O 
•— 


I 


1 ! «^-^c 





d 

} 


-4 S L_v 




d c-i b' 

Volume EntroDy 

Fig. 187. The Rankine Cycle with Incomplete Expansion (Saturated Steam). 

The work done by the piston on the steam during exhaust is 

W 3 = p 2 (x c u c + <j) foot-pounds. (108) 

The total work done is 

W t = W l + W 2 - Wi. 
Combining above equations and reducing, 
W t = (xipi + qi— x c p c — q e )+ 778 + P1X1W1 — p c x c u c -f- (pi — p 2 )(r. (109) 
The last term is small and may be omitted. 

7) X U 

Adding and subtracting c c and dividing by 778 equation (109) 
reduces to H t = H 1 -H c + (p c - p 2 ) x c u c - 778. (110) 

The steam consumption or water rate of the perfect engine operating 
in the Rankine cycle with incomplete expansion is 

W'= <2 ^ , fill) 

yVt H 1 -H c + (p c - p 2 )x c u c -r- 778' K J 

and the efficiency is 

„ , Hi — H c + (p c — p 2 )x c u c "-s- 778 n - ON 

& r = jj {HZ) 

ni — qi 



STEAM ENGINES 



315 




£ ©« 



316 



STEAM POWER PLANT ENGINEERING 




STEAM ENGINES 



317 



If p c is made equal to p 2 equation (112) will be reduced to the same 
form as equation (99) because the cycle in such case becomes complete. 
Example: Find the efficiency and the water rate of a perfect engine 
working in the incomplete cycle, using the data of the previous example, 
but assuming release to take place at a pressure of one pound above 
condenser pressure. 

Hi = 1332, q 2 = 48, as previously determined, 

Pc = P2 + 1 = 0.5 + 1 = 1.5. 

From the Mollier diagram 

H c = 965, x c = 0.856. 

From steam tables u c = 226. 
Substituting these values in (112) 

1332 - 965 -f (1.5 - 1) 0.856 X 226 ■*■ 778 
r ~ 1332 - 48 

= 0.286 or 28.6 per cent. 

176. Conventional Ideal Engine. — In designing piston engines it is 
customary to assume as a basis of reference an ideal cycle which con- 
siders only the kinetic action of the 
steam in the cylinder. This permits 
of analysis without the use of steam 
tables. The ideal diagram recom- 
mended in this connection represents 
the maximum power obtainable from 
steam accounted for by the indicator 
diagram at the point of cut-off.* 
Such a diagram for a simple non- 
condensing engine is illustrated in 
Fig. 190. AB represents admission at pressure p 1} BC represents 
hyperbolic expansion from cut-off B to release at C and DE represents 
exhaust at atmospheric pressure p 2 . The work done is represented by 
the area 

ABCDE = OABG + GBCF - OEDF. (113) 

OABG = pivi, 
GBCF 





A 




—Vi >- 


B 








/ 


. 




b 




1 


) 


v 












E 


Vl 




I 


"^TS 




Pi 





<- a 




+ 

















Fig. 190. 



Conventional Indicator 
Diagram. 



Area 



J^2 
pdv, 



OEDF = p 2 v 2 . 
Substituting these values in (113) and reducing 

Area ABCDE = Wl + log e ^- 2 ) 



P2V 2 . 



(114) 



* See Trans. A.S.M.E., vol. 24, p. 751. 



318 STEAM POWER PLANT ENGINEERING 

Letting — = r ( = ratio of expansion), 

Area ABODE = p x v x (1 + log e r) - p 2 v 2 . (115) 

The mean effective pressure = 



area ABODE 
M.e.p. = 

V 2 
Pi 

r 



= & (1 + log.*-) - p 2 , (116) 



The theoretical maximum horse power is 

PLAN 
HP - ^3,000"' < 117 > 

in which 

P = mean effective pressure, pounds per square inch, 

L = length of stroke in feet, 

A = area of cylinder in square inches, 

N = number of working strokes per minute. 

The ratio of the m.e.p. of the actual diagram abcdef to that of the 
ideal diagram as determined above is called the diagram factor. This 
factor is determined by experiment and ranges as follows: ("Heat 
Power Engineering," Hirshfeld and Barnard, 1912, p. 325.) 

Simple slide-valve engine 55 to 90 

Simple Corliss engine 85 to 90 

Compound slide-valve engine 55 to 80 

Compound Corliss engine 75 to 85 

Triple expansion engine 55 to 70 

Example: Determine the probable horse power of a 12-inch X 12- 
inch simple engine, 250 r.p.m., initial pressure 120 pounds absolute, 
cut-off J stroke, diagram factor, 0.75. 

V2 1 , 

r = — = T = 4. 

»i i 

Probable m.e.p. = 0.75 j ^ (1 + loge 4) - 15) I 

= 0.75 (30 X 2.386 - 15) 

= 42.4. 

, , , . , 42.4 X 1 X 113 X 500 
Probable i.h.p = ^^ — 

= 72.4. 



STEAM ENGINES 



319 



177. The Actual Engine. — To realize the ideal Rankine cycle the 
walls of the cylinder and the piston must be non-condensing, expansion 
after cut-off must be adiabatic, the action of the valves must be in- 
stantaneous and the steam passages must be sufficiently large to pre- 
vent wiredrawing. None of these conditions is fulfilled by the actual 
engine- The difference between the action of saturated steam in a 
perfect engine working in the Rankine cycle and that of a simple non- 
condensing engine for the same initial conditions is shown in Fig. 191 
(A) and (B). 

The area ABCD, Fig. 191 (A), represents the foot-pounds of energy 
developed per stroke by the ideal engine, and the shaded area abed the 
energy developed per stroke by the actual engine using the same weight 
of steam. The difference between the two areas represents the foot- 




A 


1' 


\ 


B 


nj 

1 lr 






i # 


6 

u 


c'\ 

10 \ 










; D 


c" 
(B) 

A' 


V 

b' 


c \ 

b' n 



. Entropy- 
Imperfections of the Actual Cycle. 

pounds of energy lost or wasted. The area ABCD, Fig. 191 (B), rep- 
resents the heat available (B.t.u. per stroke) for doing useful work in 
the ideal engine, and the shaded area abed the heat used by the actual 
engine. The difference between the two areas represents the heat lost 
or wasted. The corresponding areas in (A) and (B) are identical 
when referred to the same units. The various losses which prevent the 
actual engine from obtaining the efficiency of the ideal are outlined in 
paragraphs 187 to 195. 

178. Efficiency Standards. — The performance of a steam engine is 
variously stated as 

1. Steam consumption, pounds per hour or per horse-power hour. 

2. Heat consumption, B.t.u. per horse power per minute. 

3. Thermal efficiency, per cent. 

4. Mechanical efficiency, per cent. 

5. Efficiency ratio or potential efficiency, per cent. 

6. Cylinder efficiency. 

7. Commercial efficiency. 

8. Duty (see paragraph 309), 



320 STEAM POWER PLANT ENGINEERING 

Because of the indiscriminate use of many of these terms much con- 
fusion arises in comparing the results of different experimenters. The 
writer has consulted a number of authorities and the definitions given 
in this text are in accord with the opinion of the majority. 

The indicator offers the simplest means of measuring the output of 
a piston engine, and for this reason the performance is usually stated as 
indicated horse power. The indicated horse power is always greater than 
the net available power by an amount equivalent to the friction of the 
mechanism. The power actually developed, or brake horse power, is 
not readily obtained except for small sizes, and it is customary to ap- 
proximate this value by deducting the indicated horse power when 
running idle from the indicated horse power when running under the 
given load. This does not give the true effective power, but is suf- 
ficiently accurate for most commercial purposes. (See paragraph 192.) 
The output of steam turbines and piston engines driving electrical 
machinery is conveniently stated in electrical horse power or kilowatts, 
since the electrical measurements are readily made. The electrical 
output as measured at the switchboard gives the net effective work, 
and automatically deducts the machine losses. Large turbines are 
usually tested at the factory by means of suitable water brakes, and 
the brake horse power may be obtained from the makers. 

179. Steam Consumption. — The most generally used measure of 
the performance of a steam engine or turbine is the steam consumption 
per hour or per unit of work output. For reasons stated above the 
economy of the piston engine is given as the weight of steam consumed 
per indicated horse-power hour. This must not be confused with the 
steam accounted for by the indicator diagram, or, as it is commonly called, 
the indicated steam consumption. The former refers to the actual 
weight of fluid flowing through the cylinder and the latter to the weight 
of steam calculated from the indicator card. (See paragraph 7, 
Appendix C.) For electrically driven machinery the economy is 
given as the steam consumption per electrical horse-power hour or per 
kilowatt hour. If the initial pressure, quality and back pressure were 
constant for all conditions of operation the hourly steam consumption 
per unit output would be a true measure of the heat efficiency. (See 
Tables 59 to 66 for water rates of saturated steam engines and Tables 
69 to 73 for superheated steam engines.) 

180. Heat Consumption. — Because of the extreme variation in 
steam conditions the performance of all engines and turbines is best 
expressed in terms of the heat consumption per unit output measured 
above the maximum theoretical temperature at which the condensation 
can be returned to the boiler. This temperature is called the ideal 



STEAM ENGINES 321 

feed-water temperature. Thus the ideal feed-water temperature of an 
unjacketed non-condensing engine without receiver coils exhausting at 
standard atmospheric pressure is 212 degrees F., and that of a con- 
densing engine exhausting against an absolute back pressure of two 
pounds is 126 degrees F. If the engine is fitted with jackets and re- 
heating coils the heat of the liquid at jacket and coil pressure should 
be added to that of the exhaust in determining the ideal feed-water 
temperature. For example, if a condensing engine exhausts against an 
absolute back pressure of two pounds, and ten per cent of the total 
weight exhausted is condensed in the jackets under a pressure of 150 
pounds absolute, the ideal feed-water temperature will be 159.5 degrees 
F. (Heat of the liquid at 150 pounds absolute = 330 B.t.u. per pound. 
Heat added by the jackets to the feed water = 330 X 0.1 = 33. Heat 
of the liquid at two pounds absolute = 94 B.t.u. 94 + 33 = 127 B.t.u., 
which corresponds to an actual temperature of 159.5 degrees F.) 

Example: (1) A compound condensing engine develops one brake- 
horse-power hour on a steam consumption of 8.5 pounds, initial pressure 
200 pounds absolute, superheat 250 degrees F., exhaust pressure 0.5 pound 
absolute, release pressure two pounds absolute. (2) The same engine 
when using wet steam develops one brake-horse-power hour on a steam 
consumption of 12 pounds per hour, initial pressure 150 pounds absolute, 
quality 98 per cent, exhaust pressure two pounds absolute, release 
pressure four pounds absolute. 

Determine the comparative heat consumption of the two engines: 

Superheated steam engine, 

Hi = 1332 (from steam tables), 
q 2 = 48, 

Heat supplied per d.h.p.-hour 8.5 (1332 - 48) = 10,914 B.t.u., 
Heat supplied per d.h.p. per minute = 181.9 B.t.u., 
Saturated steam engine, 
H l = XiT! + qi 

= .98 X 863.2 + 358.5 = 1176.1, 

(This may be obtained directly from the Mollier diagram.) 
<Z2 = 94, 

Heat supplied per d.h.p.-hour = 12 (1176 - 94) = 12,985 B.t.u., 
Heat supplied per d.h.p. per minute = 216.4. 
Economy of superheated steam 

i o q 5 

(1) in steam consumption, 100 = 29.2 per cent, 

/on • u ± +• 1ftn 216.4- 181.9 tCft 

(2) in heat consumption, 100 91R . = 15.9 per cent. 



322 



STEAM POWER PLANT ENGINEERING 



181. Thermal Efficiency. — The thermal efficiency of a steam engine 
or turbine is the ratio of the heat converted into useful work to that 
supplied, measured above the heat of the liquid at exhaust pressure.* 
If the heat consumption is expressed in terms of i.h.p.-hour, the ratio 
becomes the indicated thermal efficiency. Since the heat equivalent of 



300 
275 

250 

a> 




















































/ 
































1 
















O 
,3 200 

M 

^175 

8 

j5150 

§125 

1 
£100 

a 

S 75 

CD 

50 
25 


















if 




























c 


;"/ 


J/ 


ij 
























JS-. 








w 


























V 


^ 


l7 


J 


■/ 






















c 


y 


// 




/ 








w 


















V/ 






J 








4 
J*/ 
















<» 




/ 


<*£ 










£ 


vy 


Status of the 
Piston Engine 

1907 
Saturated Steam 




i 








V 








' 


%y 





































8 10 12 14 16 18 20 22 24. 26 
Thermal Efficiency, Per Cent 

Fig. 192. 



28 30 32 34 



one horse power, using the latest accepted values, is 42.44 B.t.u. per 
minute or 2546 B.t.u. per hour, this relationship may be expressed 



E t = 



2546 



(118) 



B.t.u. supplied per b.h.p.-hour 
2546 

W(H 1 -q 2 y 
in which 

W = the weight of steam supplied, pounds per developed horse- 
power hour, 

Hi and q% as in previous equations. 

* The heat supplied is often measured above the actual feed-water temperature 
but the latter is not dependent upon the performance of the engine and hence is 
not satisfactory for purposes of comparison. The ratio of the heat converted into 
useful work to that supplied above the actual feed-water temperature is designated 
as the "thermal efficiency ratio," in the A.S.M.E. Code for Testing Steam Engines. 



STEAM ENGINES 323 

If measured in electrical units this relationship becomes 

in which 

W\ = pounds per kilowatt-hour, other notations as in (118). 

Example: Determine the thermal efficiency for the two engines using 
the data of the preceding problem. 
Superheated steam engine, 

E ' = 8.5(1332-48) = °' 233 = 23 ' 3 per Cent 

Saturated steam engine, 

2^46 
E ' = 12 (1176.1 - 94) = °- 196 = 19 - 6 per Cent 

Fig. 192 shows the present status of the piston engine, using saturated 
steam for various initial pressures. Table 66 gives the thermal efficien- 
cies for saturated steam engines, and Table 69 the thermal efficiencies 
for superheated steam engines. 

182. Mechanical Efficiency. — The ratio of the developed or brake 
horse power to the indicated power is the mechanical efficiency of the 
engine; the ratio of the electrical horse power to the indicated power is 
the mechanical efficiency of the engine and generator combined; and 
the ratio of the pump horse power to the indicated power of the engine 
is the mechanical efficiency of the engine and pumps combined. The 
percentage of work lost in friction is therefore the difference between 
100 per cent and the mechanical efficiency in per cent. (See also 
paragraph 192.) Table 56 gives the mechanical efficiency for several 
types of engines and Fig. 193 illustrates the relation between load and 
mechanical efficiency for a 75-kilowatt direct-connected engine and 
generator. 

183. Efficiency Ratio. — The degree of perfection of an engine or the 
extent to which the theoretical possibilities are realized is the ratio of 
the thermal efficiency of the actual engine to that of an ideally perfect 
engine working in the Rankine cycle with complete expansion. This 
is called the efficiency ratio or potential efficiency * It is the accepted 
standard for comparing the performance of steam engines and steam 
turbines. 

* The term "thermodynamic efficiency" or "efficiency" without qualification is 
ordinarily interpreted as the efficiency ratio, though some authorities apply the 
name "thermodynamic efficiency" to the "thermal efficiency" as defined in para- 
graph 181. 



324 



STEAM POWER PLANT ENGINEERING 



TABLE 56. 

MECHANICAL EFFICIENCIES OF ENGINES. 



Kind of Engine. 


Horse Power. 


Efficiency at Full 
Load. 


Simple: 

1. High-speed, non-condensing 


150 
170 
275 

150 

160 

900 

1000 

5500 

7500 

865 
712 


95 5 


2. High-speed, condensing 


96 


3. Low-speed, non-condensing 


94 


Compound: 

4. High-speed, non-condensing 

5. High-speed, condensing 

6. Low-speed, non-condensing 


94 
98 
95 


7. Low-speed, condensing 


95 


8. Do 


95 2* 


9. Do 


93.0* 


Triple: (combined efficiency of engine and 
pump) 

10. Pumping engine 

Quadruple: (combined efficiency of engine and 

pump) 

11. Pumping engine 


97.4 
93 



* Combined efficiency of engine and generator. 

1. Buffalo Simple engine, 12 X 12, Elec. World, Sept., 1904, p. 147. 

2. Reeves Simple engine, 15 X 14, Elec. World, Oct. 1, 1904, p. 587. 

3. 24 X 48 Hamilton Corliss at Armour Inst, of Tech., 1898. 

4. Reeves Compound, Eng. Rec, July 1, 1905, p. 24. 

5. Reeves Compound, Eng. Rec. 

6. 21, 41 X 30 Cross Compound Ball & Wood, West Albany Station, N.Y.C. & H.R.R. 

7. 20, 40 X 42 Rice and Sargent, A.S.M.E., 29-1276. 

8. N.Y. Edison, Waterside Station. 

9. New York Subway. 

10. Allis Pumping Engine, Power, May, 1906, p. 299. 



11. Nordburg Pumping Engine, Eng. News, May 4, 



p. 280. 




50 60 70 80 90 100 
Per Cent of Rated Load 

Fig. 193. 



110 150 130 140 150 



JFi(#i - 


52) 


Hi — Hi 




(Hi - q 2 ) 




2546 




Wi{Hi - 


£2) 


2546 





STEAM ENGINES 325 

If E = efficiency ratio, 

E t = thermal efficiency of the actual engine, 

E r = efficiency of the ideal engine work in the Rankine cycle with 
complete expansion, 

Then ^=l £ - (120) 

From equation (118), 

«- 2546 

And from equation (99) 

E r = 

Whence 

E = w 2546 ^ + fl.-g, 

XI 1 — #2 

T7 (ft - # 2 ) (122) 

Example: Determine the efficiency ratio of the two engines specified 
in paragraph 180. 

Superheated steam engine: 

E = 8.5 (1332 - 908) = °- 706 = 70 - 6 P er Cent ' 
Saturated steam engine: 

E = 12 (1176 - 898) = ° 763 = 76 - 3 PCT Cent - 

Tables 66 and 69 give the best recorded efficiency ratios for current 
practice. 

184. Cylinder Efficiency. — The piston engine seldom expands the 
steam down to the existing back pressure but releases from two to five 
pounds above this point in condensing engines and from 15 to 20 pounds 
above in non-condensing engines. The ideal cycle corresponding to 
this condition is the Rankine cycle with incomplete expansion. The 
ratio of the thermal efficiency of the actual engine to that of the ideal 
engine working in the incomplete cycle is a true measure of the degree 
of perfection of the engine under the given conditions. This rate is 
called cylinder efficiency and may be expressed as 

W = ?S46 

W[(Hi - H e ) + (Vc ~ P2) x c u c + 778] " K ^ 6) 

See equations (118) and (112). 



326 



STEAM POWER PLANT ENGINEERING 



E 



It may be expressed also by the relationship 

Steam consumption of the ideal engine with incomplete expansion 
Steam consumption of the actual engine 



Example: Determine the cylinder efficiency of the two engines 
specified in paragraph 180. 



Superheated steam engine, 
E' 



2546 



8.5* [1332 - 980 + (2.0 - 0.5) 0.866 X 173.5 -f- 778] 
= 0.85 = 85.5 per cent. 

Saturated steam engine, 

E' = 



2546 



12 [1176 - 935 + (4 - 2) 0.808 X 90.5 ^ 778] 
= 0.88 = 88 per cent. 

Summing up the various efficiencies for the two cases analyzed in 
paragraphs 180 to 184: 



Pressure, pounds per square inch, absolute: 

Initial 

Release 

Condenser 

Degree of superheat, degrees F 

Steam consumption, pounds per developed-horse-power 
hour 

Actual engine 

Ideal engine, Rankine cycle, with incomplete expan- 
sion 

Ideal engine, Rankine cycle, with complete expansion 

Ideal engine, Carnot cycle 

Thermal efficiency, per cent 

Actual engine 

Ideal engine, Rankine cycle, with incomplete expan- 
sion 

Ideal engine, Rankine cycle, with complete expansion. 

Ideal engine, Carnot cycle 

Heat consumption, B.t.u., per developed-horse-power 
minute 

Actual engine, Ideal engine, Rankine cycle, with in- 
complete expansion 

Ideal engine, Rankine cycle, with complete expansion. 

Ideal engine, Carnot cycle , 

Efficiency ratio, per cent 

Cylinder efficiency, per cent 



Saturated 
Steam Engine. 


Superheated. 
Steam Engine. 


150 

4 
2 

0.98* 


200 

2 

0.5 
250 


12.00 


8.50 


10.56 
9.16 
8.30 


7.22 
6.00 


19.6 


23.3 


22.3 

25.8 
28.3 


27.4 
33.3 


216.4 


181.9 


190.4 
165.1 

149.7 
76.3 

88.0 


154.6 • 

128.5 

70.6 
85.0 



* Quality. 

* If the steam consumption per i.h.p.-hour is used in this connection instead of 
the consumption per d.h.p.-hour this ratio becomes the indicated cylinder efficiency. 



STEAM ENGINES 327 

185. Commercial Efficiencies. — There is no accepted standard for 
rating the commercial efficiency of an engine or turbine. The various 
measures used in this connection, such as pounds of standard coal per 
d.h.p.-hour, cents per horse power per year and the like include the economy 
of the boiler and auxiliaries and are not a true indication of the per- 
formance of the engine alone. From a commercial standpoint it is im- 
portant to know the weight of coal required to develop a horse-power 
hour, taking into consideration all of the losses of transmission and con- 
version, and a knowledge of the over-all efficiency from switchboard 
to coal pile is of value in basing the cost of power, but these items 
are in reality measures of the plant economy and are of little value in 
comparing the performance of the prime mover. The various efficiencies 
under this heading are treated in Appendices C to G. 

186. Heat Losses in the Steam Engine. — The principal losses which 
tend to lower the efficiency of the steam engine and which prevent it 
from realizing the performance of the ideal engine are due to 

(a) Cylinder condensation. 

(6) Leakage. 

. (c) Clearance volume. 

(d) Incomplete expansion. 

(e) Wire drawing. 

(/) Friction of the mechanism. 

(g) Presence of moisture in the steam at admission. 

(h) Radiation, convection and minor losses. 

187. Cylinder Condensation. — The weight of steam apparently used 
per revolution, as determined from the indicator card, or the indicated 
steam consumption* (see paragraph 7, Appendix C) is considerably 
less than that actually supplied. The difference or missing quantity is 
due chiefly to cylinder condensation. This is by far the greatest loss in 
the steam engine with the exception of that inherent in the ideal en- 
gine. When steam is admitted to the cylinder a considerable portion of 
the heat is given up to the comparatively cool skin surface of the 
cylinder walls. If the steam is saturated at admission this heat ab- 
sorption causes condensation, or initial condensation as it is called; if 
superheated at admission the temperature is lowered to a correspond- 
ing point. After cut-off heat continues to be given up to the walls 
until' the temperature of the steam falls below that of the skin surface, 
when the process is reversed and part of the heat is returned to the 
steam. With saturated steam the heat absorption causes condensation 
during expansion and the heat rejected, reevaporation during expansion, 

* Also called the steam accounted for by the diagram or diagram steam. 



328 



STEAM POWER PLANT ENGINEERING 



With superheated steam an equivalent heat exchange takes place. 
Unless the cylinder is of a compound series the heat absorbed from the 
cylinder walls during exhaust does no useful work and is lost. Cylinder 
condensation, measured as the proportion of the mixture present, 
varies with the size of the engine, speed, length of cut-off, valve de- 
sign, temperature range, location of ports and port passages, jacket- 
ing, lagging, and other variables. It ranges from 16 to 30 per cent, 
and is occasionally as high as 50 per cent of the total weight of steam 
admitted to the cylinder. Cylinder condensation and leakage are 



CO 



50 



40 



30 



I 20 



10 





o 












































































Condensation and Leakage 
for 
Simple Engines 

using 
Saturated Steam> 














o 






















\o 
































o 


s1 






























o 





n 


sO 
































O 


\c 









































D r 






































O 






















































Ej 


lgine 


Tests 


Ban 


us, p. 


254. 



15 20 25 

Percentage of Cut Off 

Fig. 194. 



40 



ordinarily classified together for sake of simplicity. Fig. 194 shows 
the relation between cylinder condensation and leakage losses for vari- 
ous percentages of cut-off for simple high-speed non-condensing engines. 

Empirical formulas for calculating the extent of these losses, and 
which involve the various influencing factors, are unwieldly and only 
approximately accurate. One of the most satisfactory formulas of 
this class is that deduced by R. C. H. Heck, " The Steam Engine and 
Turbine," 1911, p. 175. 

The various heat exchanges between working fluid and the cylinder 
walls, including cylinder condensation and leakage, are approximately 
determined by transferring the indicator diagram to the temperature 
entropy chart. (See Fig. 191.) 



STEAM ENGINES 



329 



For use and application of the temperature-entropy diagram in engine 
tests consult Power, Dec, 1907, p. 834; Jan. 21, 1908, p. 96; Jan. 28, 
1908, p. 145. 

A comparatively simple method for approximating cylinder con- 
densation and leakage losses is given by J. Paul Clayton, Proc. A.S.M.E., 
April/ 1912, and consists in transferring the indicator diagram to loga- 
rithmic cross-section paper. By means of the logarithmic diagram 
Clayton found that (1) free from certain abnormal influences, expansion 
and compression take place in the cylinder substantially according to 
the law PV n = C, (2) the value n bears a definite relation in any given 
cylinder to the proportion of the total weight of steam mixture which 
was present as steam at cut-off, (3) the relation of the value n to the 
value X c (quality of steam at cut-off) for the same. class of cylinder as 



57.3 
58.1 

38.7 

29.1 H 

19.8 G 

14.5 F 



=£* 



131.3 
115.3 

96.3 
77.3 
58.1 

38.7 

29.2 
Q 19.8 
N 14.5 



CRANK END 



19.5 
14.5 

0.0 



Fig. 195. 





\ 




x 




V 








\ ^ — --^ 




===== " 



95.0 
76.3 
57.1 
38.9 
29.4 
19.5 
14.5 
0.0 
HEAD END 

Diagrams taken from a 12 X 24 Corliss Engine. 



regards jacketing is practically independent of engine speed and of 
cylinder size, and (4) by means of the experimentally determined re- 
lation of X n and n the actual steam consumption may be obtained 
from the indicator card to well within 4 per cent of the true value. 
The curves in Fig. 196 were plotted on logarithmic cross-section paper 
from the pressure-volume diagrams, shown in Fig. 195, and illustrate 
Mr. Clayton's method of analysis. The curves in Figs. 197 and 198 
show the relation between quality X c and exponent n or for a given 
set of conditions. For applications of this method to concrete examples 
with a full discussion of the results consult the paper referred to. 

188. Leakage of Steam. — The loss due to leakage is a variable 
factor depending upon the design and condition of the engine, and is 
greater with saturated than with superheated steam. The usual 



330 



STEAM POWER PLANT ENGINEERING 



method of measuring leakage past the valves and piston while the 
engine is at rest is likely to give erroneous results, as demonstrated by 
Callender and Nicolson (Peabody, "Thermodynamics/' p. 351) in tests 



f 


r 


.0 — 




CfRjS 


X 

Hip— 














— -^ 















— -*- 




CRX 


H 






\ 














X 
























w 




















'::z..=j r 








iff B 







0.1 



2.0 



0.2 0.3 0.4 0.5 0.6 0.70.80.91.0 

Absolute Volume - Cu. Ft. 
Fig. 196. Logarithmic Diagrams plotted from Fig. 195. 

made on a high-speed automatic balanced valve engine and on a quad- 
ruple expansion engine with plain unbalanced slide valves. With the 
engines at rest they found that the leakage past valves and piston was 



90 - 


















j 














i 80- 
































70 - 
















































:& — a-o = 
















d 


















"2iJ 
































T 5 


















































\\ 




























































© 






















































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.P 

t-3 20- 






















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• 

2 


























Sv 






3 
































* *R 
































































■5 


























/ 






° 6- 


























/ 






ss _ 


























,/ 






*5 R. 


























/ 






< 5 
























/ 




























'•-. 




y 








































3. 
































0.07 


.OS 

08 I 


.1 


0. 


2 





3 





4 0.5 0. 


,0.7 0.9 
' 0.8 1 


.0 


2 





3 





1.0 



Fig. 197. 



Absolute Volume - Cu. Ft. 
Diagram from a 14 X 35 Corliss Engine, showing Leakage at Beginning and 
End of Expansion and Compression. 



insignificant, but when in operation the leakage from the steam chest 
into the exhaust was very considerable indeed. It was thought that a 



STEAM ENGINES 



331 



large proportion of the leakage was probably in the form of water 
formed by condensation of steam on the seat uncovered by the valve. 

According to the report of the Steam Engine Research Committee 
(Eng. Lond., March 24, 1905, p. 208), leakage through a plain slide 
valve is independent of the speed of the sliding surfaces, and directly 
proportional to the difference in pressure on the two sides; with well- 
fitted valves the leakage is never less than 4 per cent of the volume of 
steam entering the cylinders, and is often greater than 20 per cent. 

The various leakage losses may be approximated by transferring the 
indicator diagram to logarithmic cross-section paper. Fig. 197 shows 
the application of the logarithmic diagram to a specific case and illus- 
trates this method of determining leakage losses. See Clayton's paper 
referred to in paragraph 188. 

Leakage Past Piston Valves: Engr., Feb. 9, 1912. 



X00 
0.95 

0.90 
0.85 
0.80 
0.75 
0.70 
0.65 
0.60 
0.55 
0.50 
0.45 

0.850 0.900 0.950 1.000 1.050 1.100 1.150 1.200 1.250 
Value of n f rom Expansion Curve 

Fig. 198. Relation of Quality and the Value of n. 

189. Clearance Volume. — The portion of the cylinder volume not 
swept through by the piston but which is nevertheless filled with steam 
when admission occurs is called the clearance volume. It is the space 
between the end of the piston when on dead center and the inside of 
the valves covering the ports. It varies from about 1 per cent of the 
piston displacement in very large engines with short steam passages 
to 10 per cent or more in small high-speed engines. When the steam 
retained in the clearance space is compressed to the initial pressure 
and expansion is carried down to the back pressure, the clearance has 
little effect upon the economy of the engine, but since expansion and 
compression are seldom complete in actual practice, the loss may be con- 
siderable. (Ripper, "Steam Engine," p. 103.) The shorter the cut-off 





































o 


































2 
o 




























s 


y 




ta 






















































































° 
















S 


















^° 
















<5 














/» 




















"41 

o 








o^ 




























































3 


































U* 




















o Test with saturated Steam 
a Test with Superheated Steam 
























1 

Efc 


KJent 
uatio 


er org 
n.ofc 


nditi 


mXp 


pom 
1.258 


M-0. 


>14 



332 



STEAM POWER PLANT ENGINEERING 



the greater will be the ratio of the weight of cushion steam to that of 
the steam supplied and hence the greater the relative loss. In large slow- 
speed engines the loss may be insignificant if the clearance volumes are 
small, while in small high-speed engines it may be considerable. 

The ratio of expansion is decreased by clearance; for example, an 
engine cutting off at one-fifth, neglecting clearance, has an apparent 
ratio of expansion of 5, but if the clearance volume is 10 per cent the 
actual ratio is only 3.66. One of the few recorded tests relative to the 
influence of clearance on the economy of a high-speed engine was con- 
ducted on a 14 X 15 Allfree engine. (Power, May, 1901.) With a 
clearance volume of 2.2 per cent, initial pressure 105 pounds gauge, 
and 172 r.p.m., the best performance was 23.7 pounds of dry steam 
per i.h.p. hour. With the same steam pressure and speed, but with 
clearance volume increased to 6 per cent by the use of a shorter piston, 
the best performance was 28.3 pounds per i.h.p. hour. In both cases 
the compression was carried up to admission pressure. 

Independent tests made by Prof. Boulvin and by A. H. Klemperer 
on single-cylinder Corliss engines gave a minimum water rate when the 
clearance volume was approximately one-half the compression volume. 
See end of paragraph 190. 

Engine Clearance and Compression: Power, July 5, 1910, Dec. 27, 1910; Sibley 
Journal, Dec., 1910. 

190. Loss due to Incomplete Expansion and Compression. — Theo- 
retically the loss due to incomplete expansion is considerable. For 
example, the theoretical steam consumption of a perfect engine (Ran- 
kine cycle) expanding from 120 pounds absolute to a condenser pres- 
sure of 2 pounds absolute is 9.6 pounds per horse-power hour. If the 
expansion were carried to only 5 pounds absolute, the exhaust pressure 
remaining the same, the steam consumption would be increased to 
11.8 pounds per horse-power hour, a difference of 22 per cent for an 
increase in terminal pressure of only 3 pounds per square inch. The 
theoretical water rates for various terminal pressures are given below. 



Terminal Pressure, 

Pounds per Square 

Inch Absolute. 


Steam Consumption 
of Perfect Engine. 


Terminal Pressure, 

Pounds per Square 

Inch Absolute. 


Steam Consumption 
of Perfect Engine. 


1 

1.5 

2 

2.5 


8.5 
9.1 
9.6 

10 


3 
4 
5 
6 


10.4 
11.1 
11.8 
12.3 



In actual engines expansion is seldom complete, since it would 
necessitate increased bulk and weight of engine, and the work done 



STEAM ENGINES 



333 



by the steam in the last stages would not compensate for the increased 
cost. 

In single-cylinder engines maximum economy is effected when the 
terminal pressure is considerably above that of the exhaust, since the 
gain due to complete expansion is more than offset by the increased 
cylinder condensation. This is true to a certain extent in all engines 
irrespective of the number of cylinders. Tests by G. H. Barrus 
(" Engine Tests/' 1900) to determine the terminal pressures effecting 
maximum economy for various types of engine gave results as follows: 



Simple slide-valve engines, non-condensing 
Simple slide-valve engines, condensing .... 
Simple Corliss engines, non-condensing. . . . 

Simple Corliss engines, condensing 

Compound engines, non-condensing 

Compound engines, condensing 



Terminal Pressure, 


Pounds Absolute. 


30 to 40 


25 to-30 


20 to 25 


15 to 18 


18 to 22 


3 to 5 



In high-speed engines a certain amount of compression is desirable 
for its cushioning effect; outside of this mechanical feature compres- 
sion may or may not be of benefit to the engine, as will be seen from 
the results of tests stated below. Zeuner in his treatise on theoreti- 
cal thermodynamics proves deductively that in an engine with a large 
clearance volume the loss due to clearance is completely eliminated if 
the compression is carried up to admission pressure, a conclusion which 
tests by Jacobus, Carpenter, and others fail to confirm. A series of 
tests by Professor Jacobus (Trans. A.S.M.E., 15-918) on a 10 X 11 
high-speed automatic engine at Stevens Institute show decreasing 
economy with increase of compression, the initial pressure, cut-off, and 
release remaining constant. The results were as follows: 



Proportion of initial pressure up to which the 

steam is compressed 

Steam, pounds per i.h.p. hour 



5 


2 


8 


3 


34.8 


36.7 



Full 
38 



Tests by Carpenter (Trans. A.S.M.E., 16-957) on the high-pressure 
cylinders of the Corliss engine at Sibley College gave: 



Compression, per cent 

Brake horse power 

Steam, pounds per b.h.p. hour. 



11.4 


25 


30 


29 


33 


33.3 



35.2 

26 

34 



Tests, made by A. H. Klemperer on a 7.1 X 17.7-inch Corliss engine, 
at Dresden, gave decreasing steam consumption for increase in com- 



334 



STEAM POWER PLANT ENGINEERING 



pression up to a compression of about twice the clearance volume, be- 
yond which the water rate increased with the increase in compression. 
(Zeit d. Ver. deut. Ingr., Vol. I, 1905, p. 797.) 

Tests, made by Prof. Boulvin on a 9.8 X 19.7-inch Corliss engine at 
University of Ghent, give results agreeing with those of Klemperer. 
(Revue de Mecanique, 1907, Vol. XX, p. 109.) 

Fig. 199 shows the influence of increasing back pressure on the 
economy of an 8 X 10-inch automatic high-speed engine at the Armour 
Institute of Technology. 



6.8 
6.7 > 
6.6 

<x> 

O 6 4 














. 








43 
47 
46 
45 
44 














100 Lb. Gauge 






































^s^* 
















« 










h 

S> 

* 6<3 

| 6.2 

o 

1 6.1 

■a 6 -° 

| 5.9 

^ 5.8' 

5.7 

5.6 

5.5 






\ 


^ 




^ 


ps 




























1° 










^^ 












41 
10 






^ 




<fy 


C Vs 




















> 










39 
38 
37 
,36 


^ 




Influence of Back 

Pressure on the Economy of 

an 8 x 10 Automatic High Speed 

STon Condensing Engine 














































^> 



6 8 10 12 14 

Back Pressure.L'b. Per Sq.In.Gauge 
Fig. 199. 



16 



18 



20 



191. Loss due to Wire Drawing. — Wire drawing, or the drop in 
pressure due to the resistances of the ports and passages, has the effect 
of reducing the output and the economy of the engine to some extent, 
since the pressure within the cylinder is less than that at the throttle 
during admission and greater than discharge pressure at exhaust. 
The steam may be dried to a small extent during admission, but be- 
cause of the drop in pressure the heat availability is reduced. In single- 
valve engines the effects of wire drawing are decidedly marked and the 
true points of cut-off and release are sometimes difficult to locate on 
the indicator card. In engines of the Corliss or gridiron- valve type 
the effects are hardly noticeable. 

192. Loss due to Friction of the Mechanism. — The difference between 
the indicated horse power and that actually developed is the power 
consumed in overcoming friction, and varies from 4 to 20 per cent of 
the indicated power, depending upon the type and condition of the 



STEAM ENGINES 



335 



engine. Engine friction may be divided into (1) initial or no-load 
friction and (2) load friction. The stuffing-box and piston-ring friction 
is practically independent of the load, while that of the guides, bear- 




25 50 75 100 125 150 175 200 225 250 275 300 

IndicatedJSorse-Power 
Fig. 200. Typical Curves of Steam Engine Friction. 



ings, and the like increases with the load. In Fig. 200, curve A gives 
the relation between the frictions for a four-slide-valve horizontal 
cross compound engine, and B that for a simple non-condensing Corliss. 
(Peabody's " Thermodynamics," pp. 433 and 437.) Curve C is plotted 

TABLE 57. 
DISTRIBUTION OF FRICTION IN SOME DIRECT-ACTING STEAM ENGINES. 

(Thurston.)* 





Percentage of Total Engine Friction. 


Parts of Engines where Friction 
is Measured. 


" Straight 
Line " 
Balanced 
Valve. 


" Straight 
Line " 
Unbalanced 
Valve. 


Traction 
Engine 
Locomotive 
Valve Gear. 


Automatic 

Balanced 

Valve. 


Condensing 
Engine 
Balanced 
Valve. 


Main bearings 


47.0 


35.4 


35.0 


41.6 


46.0 






Piston and piston rod 


•32.9 


25.0 


21.0 


49.1 




Crank pin 


6.8 


5.1 


13.0 


21.8 


Crosshead and wrist pin 


5.4 


4.1 




Valve and valve rod 


2.5 


26.4 


22.0 


9.3 






21.0 


Eccentric strap 


5.4 


4.0 


Link and eccentric 






9.0 
















Air pump 










12.0 
















100.0 


100.0 


100.0 


100.0 


100.0 



* " Friction and Lost Work in Machinery," p. 13. 



336 STEAM POWER PLANT ENGINEERING 

from the tests of a Reeves vertical cross compound condensing engine 
(Engineering Record, July 1, 1905, p. 24), and D from the test of an 
Ames simple high-speed non-condensing engine. (Engineering Record, 
Vol. 27, p. 225.) A large number of recorded tests show less friction at 
full load than at no load, but this is probably due to error or to varia-. 
tions in lubrication. With first-class lubrication it is usually sufficiently 
accurate to assume the friction to be constant and equal to the initial 
friction at zero load. The distribution of the frictional losses in a 
number of engines is given in Table 57. 

193. Moisture. — The presence of moisture in the steam pipe is due 
to condensation caused by radiation or to priming at the boiler. Un- 
less removed by some separating device between boiler and engine 
the amount of moisture entering the cylinder may be from 1 to 5 per 
cent of the total weight of steam, and the work done per pound of 
fluid is correspondingly reduced. This loss should not be charged 
against the engine, however, and its performance should be reckoned 
on the dry steam basis. Experiments reported by Professor R. C. Car- 
penter (Trans. A.S.M.E., 15-438) in which water in varying quantities 
was introduced into the steam pipe, causing the quality of the steam 
to range from 99 per cent to 57 per cent, showed that the consump- 
tion of dry steam per i.h.p. hour was practically constant, the water 
acting as an inert quantity. An efficient separator will remove prac- 
tically all the entrained water. 

194. Radiation and Minor Losses. — The radiation and conduction of 
heat from the cylinder, piston rod and valve stem has the effect of in- 
creasing the cylinder condensation. In jacket engines this loss may be 
approximated by the quantity of steam condensed in the jacket when 
the engine is not running. In unjacketed engines the loss is practically 
underterminable since the heat exchange between cylinder walls and 
the steam is exceedingly complex. 

195. Heat Lost in the Exhaust. — Most of the heat supplied to the 
engine is rejected to the exhaust; this varies from 65 per cent in the best 
types of engines to 95 per cent in the poorer types. If all of the exhaust 
is used for heating or manufacturing purposes the heat chargeable to 
power is the difference between the heat supplied and that rejected. 
In determining the latter it is necessary to know the quality of the 
exhaust steam since a considerable portion of the fluid discharged is 
water. The quality varies from 92 per cent in high-speed non-con- 
densing engines running at full load to 80 per cent or lower in compound 
non-condensing engines operating at light loads. For example, a 
24 X 18-inch simple engine direct connected to a 200-kilowatt gener- 
ator, installed at the Armour Institute of Technology, uses 55 pounds 



STEAM ENGINES 337 

of steam per kilowatt hour at full load, initial pressure 115 pounds 
absolute, back pressure 2 pounds gauge. The quality of the exhaust 
entering the heating system is 90 per cent. During the summer months 
when the exhaust is discharged to waste the entire heat supplied above 
the feed-water temperature of 210 degrees F. is chargeable to power, 
an extravagant waste of heat. During the winter months when all of 
the exhaust is used for heating purposes the heat chargeable to power 
is 55 [1188.8 - (0.9 X 965.6 + 187.5)] = 7276 B.t u. per kilowatt-hour, 
which is equivalent to 7276 -f- [1188 - (210 - 32)] = 7.2 pounds of 
boiler steam per kilowatt-hour, a performance unequalled by any com- 
pound condensing engine. With condensing engines, in which no dis- 
position is made of the heat absorbed by the circulating water, which 
is the usual case, all of the heat rejected to the exhaust less the small 
amount reclaimed from the hot well is chargeable to power, The 
latent heat rejected to exhaust, however, is an inherent loss, even for 
the ideal engine operating in the Rankine cycle. 

196. Methods for Increasing Economy. — Various methods have been 
adopted for bettering the economy of piston engines, among them may 
be mentioned: 

(a) Increasing boiler pressure. 
(6) Use of receiver reheaters. 

(c) Steam jackets. 

(d) Increasing rotative speed. 
■(e) Compounding. 

(/) Superheating. 

(g) Decreasing back pressure. 

(h) Use of binary vapors. 

(i) Use of uniflow or straight flow cylinders. 

Some of these items will be considered separately, others will be in- 
cluded in the discussion of the different classes of engines. 

197. Effect of Increased Steam Pressure. — A consideration of the 
Rankine and Carnot Cycles indicates that theoretically the greater the 
temperature range the greater will be the efficiency. (See Fig. 192.) 
In the actual engine the temperature range is most readily increased 
by raising the boiler pressure, since the limit of the back pressure is 
practically fixed by the cooling medium in the condenser. The theoreti- 
cal gain resulting from increased pressure range is, however, very con- 
siderably affected by the increased losses due to cylinder condensation. 

Fig. 201 shows the results of tests made at the Armour Institute of 
Technology on an 8 X 10 automatic high-speed piston- valve engine, 
showing marked gain with increase of initial pressure up to a certain 



338 



STEAM POWER PLANT ENGINEERING 



point when the condensation losses became sufficiently great to neu- 
tralize the advantage which would otherwise be gained. 

The following figures were obtained in tests of a small Willans engine, 
non-condensing, under different steam pressures: 



Initial Pressure, Gauge. 


Pounds Steam per I.H.P. 


B.T.U. per I.H.P. per 




Hour. 


Minute. 


36.3 


42.8 


700 


51.0 


36.0 


595 


74.0 


32.6 


544 


85.0 


29.7 


495 


97.0 


26.9 


450 


110.0 


27.8 


465 


122.0 


26.0 


436 



Referring to Fig. 192, it may be noted that both the theoretical and 
the actual efficiencies increase very slowly for pressures above 150 
pounds. Practically, gain in efficiency due to increasing the pressure 



6.2 



6.1 



£6.0 

f 
.§5.9 

e 

3 5.* 

a 



.a 
^5.7 



5.6 



5.5 



















_ 




\ x 












• ^^ 






\ 


N 




















Na 


• S 
















J* 


5s. 














4 


^ 


% 


c 'nS> 






- 




/ 


/ 














• 


/ 












> 


2— 



47 9 



-45 3 



44 



75 



90 95 100 105 

Initial Gauge Pressure.Lb.per Sq.In, 



110 



120 



42 



Fig. 201. Influence of Initial Pressure on the Economy of a Small, High-speed, Non- 
condensing Engine. 



above about 200 pounds is at the expense of increased first cost and 
maintenance and is only resorted to when small weight and space are 
the most important considerations. 



STEAM ENGINES 



339 



The range of pressures sanctioned by modern practice for different 
types of engines is as follows: 



Type of Engine. 



Simple* slow-speed (standard type) 
Simple high-speed (standard type) 

Simple, uniflow 

Compound high-speed, non-condensing 
Compound high-speed, condensing. . . . 
Compound slow-speed, condensing. . . . 

Triple expansion, condensing 

Quadruple expansion, condensing 



Range in Pressure 
(Gauge). 


Average. 


60-120 


90 


70-125 


100 


125-225 


175 


100-180 


140 


100-180 


140 


125-200 


170 


140-250 


200 


125-275 


225 



198. Receiver reheaters: Intermediate Reheating. — The receivers be- 
tween the cylinders of multi-expansion engines are frequently equipped 
with heating coils, as illustrated in Fig. 447, the function of which is to 
superheat the exhaust steam before delivering it to the cylinder im- 
mediately following, with a view of reducing the losses occasioned by 
cylinder condensation. The coils are supplied with live steam under 
boiler pressure and may serve to evaporate a portion of the moisture 
or to actually superheat the steam supplied to the following cylinder. 
The question of the propriety of using reheaters is an open one, since 
reliable data relative to their use are meager and discordant. The con- 
ditions under which the few recorded tests were made are too diverse 
to warrant definite conclusions. Some show an appreciable gain in 
economy, others a decided loss. A reheater is of little value in improv- 
ing the thermodynamic action of the engine, and is probably a loss 
unless it produces a superheat of at least 30 degrees F., and to be fully 
effective should superheat above 100 degrees F. (L. S. Marks, Trans. 
A.S.M.E., 25-500.) The effectiveness of the reheater will evidently be 
increased by the removal of the greater portion of the moisture from 
the exhaust steam before it enters the receiver. In the 5000-horse- 
power engine at the Waterside Station in New York it was shown that 
both jackets and reheaters, either together or alone, were practically 
valueless throughout the working range of load. (Power, July, 1904, 
p. 424.) Many similar cases may be cited which show no gain in 
economy with the use of the reheaters. In all cases the reheater effects 
a great reduction in the condensation in the low-pressure cylinders, 
but the resulting gain, considering the condensation in the reheater 
coils, may be little if any. On the other hand, with properly propor- 
tioned reheaters, the gain may be considerable and particularly with 
superheated steam. Practically all European engines operating with 
highly superheated steam are equipped with receiver-reheaters. In 



340 STEAM POWER PLANT ENGINEERING 

the locomobile type of engine plant the intermediate reheating is 
effected by heating coils placed in the path of the furnace gases. For a 
complete description of the locomobile with results of tests under vary- 
ing conditions of operations, see Zeit. des. Ver. deut. Ingr., Vol. 55, 
1911, p. 410 (serial); ibid, p. 922. See also Fig. 226. 

In triple-expansion pumping engines receiver- reheaters are found to 
effect an appreciable gain in economy, and practically all such engines 
are equipped with them. In electric traction work or where the load 
is a widely fluctuating one the reheater has been virtualjy abandoned. 
Apart from the consideration of fuel economy, all tests show a marked 
increase in the indicated power of the low-pressure cylinder (5 to 15 
per cent), and to that extent it increases the capacity of the entire 
engine. (G. H. Barrus, Power, Sept., 1903, p. 516.) 

Engine Reheaters: Mech. Engr., Dec. 23. 1910. 

199. Jackets. — If the walls of the cylinder are made double and 
the space between is filled with live steam under boiler pressure, the 
cylinder is said to be steam jacketed. The function of the jacket is 
to reduce initial condensation by maintaining the temperature of the 
internal walls as nearly as possible equal to that of the entering steam. 
The heat given up by the jacket steam, and the resulting condensa- 
tion, is usually a smaller loss than would otherwise result from cylinder 
condensation. However, tests of numerous engines with and without 
steam jackets do not agree as to the conditions under which their 
use is profitable, the apparent gain ranging from zero to 30 per cent. 
According to Peabody, a saving of from 5 to 10 per cent may be made by 
jacketing simple and compound condensing engines, and a saving of 
from 10 to 15 per cent by jacketing triple expansion engines of 300 horse 
power and under. On large engines of 1000 horse power or more the 
gain, if any, is very small. (Peabody, " Thermodynamics," p. 400.) 

Other things being equal, the smaller the cylinder and the lower the 
piston speed the greater is the value of the jacket. Experiments 
show no advantage in increasing the jacket pressure more than a few 
pounds above that of the initial steam in the cylinder, and it is usual 
to reduce the pressure in the jackets of the second and succeeding 
cylinders of multi-expansion engines. (Ripper, " Steam Engine/' p. 170.) 

To be effective, jackets should be well drained, kept full of live steam, 
and the water of condensation returned directly to the boiler. 

Pumping engines and other slow-speed engines running at practi- 
cally constant load are generally jacketed, but in street-railway work 
and in the majority of manufacturing plants carrying fluctuating load, 
jackets are not considered advantageous. 



STEAM ENGINES 341 

Whatever may be the actual economy due to jacketing, there is no 
question but that the jacket greatly influences the action of the steam 
in the cylinders, and whether beneficially or not depends upon the 
design and construction of the engine. Unless otherwise specified, 
manufacturers usually build their engines without jackets. 

200. .Increasing Rotative Speed. — High rotative speed does not 
necessarily mean high piston speed. An 8 X 10 engine running at 
300 r.p.m. has a piston speed of only 500 feet per minute, whereas a 
36 X 72 Corliss running at 60 r.p.m. has a piston speed of 720 feet per 
minute. The classification "high speed" and "low speed" refers to 
rotative speed only, the former above and the latter below, say 150 
r.p.m. 

On account of the reduction of thermodynamic wastes, a high-speed 
engine should give theoretically a higher efficiency than the same en- 
gine at a lower speed, all other conditions being the same. The effect 
of speed upon economy is decidedly marked in engines and pumps 
taking steam full stroke. For example, tests of a 12 X 7i X 12 simplex 
direct-acting steam pump at Armour Institute of Technology showed 
a steam consumption of 300 pounds per i.h.p. hour at 10 strokes per 
minute, and only 99 pounds at 100 strokes per minute. (See Figs. 383 
and 384.) 

Tests of engines using steam expansively, however, do not furnish 
conclusive evidence on this point, some showing a decided gain (Pea- 
body, "Thermodynamics," p. 425), others little or no gain (Barrus, 
"Engine Tests," p. 260). For example, a small Willans engine showed 
an increase in economy of 20 per cent in increasing the rotative speed 
from 111 to 408 r.p.m. (Peabody, "Thermodynamics," p. 402), whereas 
the compound locomotive at the Louisiana Purchase Exposition showed 
a loss in economy for the higher speeds (Publication by the Pennsyl- 
vania Railroad Company). On the other hand, a comparison of the 
performances of high- and low-speed Corliss engines shows little differ- 
ence in economy, and a general comparison between high- and low-speed 
engines furnishes little information, since nearly all high-speed engines 
are of a different class from the low-speed ones. High-speed engines are 
comparatively small in size, require larger clearance volume, and are 
usually fitted with a single valve. Rotative speed is limited by design, 
material, workmanship, and cost of subsequent maintenance. Speeds 
of 400 r.p.m. and more are not unusual with single-acting engines, 
whereas 300 r.p.m. is about the limit for double-acting machines with 
strokes over 12 inches in length. A comparison of tests of high-speed 
and low-speed engines in this country, irrespective of design and con- 
struction, shows the former to be less economical than the latter in 



342 STEAM POWER PLANT ENGINEERING 

most cases. In Europe high-speed engines are developed to a high 
degree of efficiency, and their performances are comparable with the 
best grade of low-speed engines. 

High-speed engines as a class have the advantage of being more 
compact for a given power, are simple in construction and relatively 
low in first cost; on the other hand, they are subject to comparatively 
rapid depreciation, excessive vibration, and are less economical in 
steam consumption. 

201. High-speed Single- valve Simple Engines. — This style of engine 
is made in sizes varying from 10 to 500 horse power. The cylinder 
dimensions vary from 4 X 5 to 24 X 24 and the rotative speed from 300 
to 175 r.p.m. 

When ground is limited or costly and exhaust steam is necessary 
for heating or manufacturing purposes, the high-speed non-condensing 
engine is most suitable for horse powers of 200 or less, being compact, 
simple in construction and operation, and low in first cost. For sizes 
larger than this the compound engine would probably prove a better 
investment, except in cases where fuel is very cheap or large quantities 
of exhaust steam are to be used for manufacturing purposes. 

Small high-speed engines are seldom operated condensing, since the 
gain due to reduction of back pressure is more than offset by the extra 
cost of the condenser and appurtenances. 

Engines are ordinarily rated at about 75 per cent of their maximum 
output. For example, a 12 X 12 non-condensing engine running at 
300 r.p.m., with initial steam pressure of 80 pounds gauge, is normally 
rated at 70 horse power, though it is capable of developing 90 horse 
power at the same speed. 

The steam consumption of high-speed single-valve non-condensing 
engines at full load ranges from 27 to 50 pounds per indicated horse- 
power hour, depending upon the size of the unit and the conditions of 
operation. An average for good practice is not far from 30 pounds. 
With superheated steam a steam consumption as low as 18 pounds 
per horse-power hour has been recorded. 

Table 59 gives the steam consumption of a number of single-valve 
high-speed engines running condensing and non-condensing, and 
Fig. 202 shows some of the results for different loads. The steam 
consumption is fairly constant from 50 per cent of the rated load to 25 
per cent overload, but for earlier loads the economy drops off rapidly. 
The desirability of operating the engine near its rated load is at once 
apparent. The curves show a marked economy in favor of the larger 
cylinders, but the engines are not of the same make, and the conditions 
of operation are somewhat different. 



STEAM ENGINES 



343 



The most economical cut-off for a simple engine is about one-third 
to one-fourth stroke when running non-condensing, and about one- 
sixth when running condensing. 

The performances given in Table 59 are exceptional. It is not ad- 
visable to count on a better steam consumption for this type of engine 
than 30 to 35 pounds of steam per i.h.p. hour. 

Fig. 206 shows the effects of condensing on a typical single-valve 
high-speed engine. The gain in fuel economy may be only an apparent 
one, since the steam consumption of the condensing apparatus should 
be rightfully charged to the engine. 



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Per Cent of Rated Load 

Fig. 202. Typical Economy Curves of High-speed, Single-valve, Non-condensing Engines. 

Saturated Steam. 

Fig. 207 shows the relation between total steam supplied, unit water 
rate and quality of exhaust at various loads for a 22 X 18 high-speed, 
single-valve, non-condensing engine direct connected to a 200-kilowatt 
direct-current generator as installed in the power plant of the Armour 
Institute of Technology. 

These curves afford a means of determining the heat actually required 
to furnish power when the exhaust is used for heating purposes. During 
the summer months the total heat supplied measured above the feed- 
water temperature is chargeable to power. During the winter months 
the heat required to furnish the electrical energy is the difference be- 
tween the total heat supplied and that exhausted to the heating system. 
(See paragraph 194.) This latter is readily obtained from the quality 



344 



STEAM POWER PLANT ENGINEERING 




Fig. 203.. Assembly of Valve Gear; Typical Corliss Engine. 



Back Cylinder Head 
Back Cylinder Head Studs 
Back Cyl. Head Bonnet 



Corliss Exhaust Valve 



Steam pipe 
Steam Flange 

Throttle Valve 
Planished Steel Lagging 
Heat Insulating Filling 

Corliss Steam Valve Chamber 
Front Cylinder Head 




Front Cylinder Head Studs 

^k^Piston Rod Gland Studs 
■/ 
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Piston Rod Packing 



Corliss Exhaust Valve 
-s — Planished Sheet Steel Lagging; 



Heat Insulating Filling , 



^Exhaust Flange 
^Exhaust Opening 
Exhaust Pipe 

Fig. 204. Section Through Cylinder; Typical Corliss Engine. 



STEAM ENGINES 



345 




346 



STEAM POWER PLANT ENGINEERING 



curve. 



By means of an indicating or recording flow meter placed in 
the exhaust line leading to the heating system the heat chargeable to 
power may be readily obtained during the months when only part of 



70 



60 



TEST OF REEVES SIMPLE ENGINE 

MADE BY PROF. R. C. CARPENTER AND PROF. H. DIEDERICH8 



AT SIBLEY COLLEGE, ^CORNELL UNIVERSITY, ITHACA, N. Y., AUGUST T804- 




100 120 

Fig. 206. 



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I 

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160 



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120 



100 



it 



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207. — Performance of 200-kilowatt Direct-connected Engine-generator Set. 
Single, Non-condensing Engine. 



the exhaust is used for heating purposes. Although the quality curves 
refer to a special case they may be used as a means of roughly approxi- 
mating the heat exhausted by any high-speed non-condensing engine. 



STEAM ENGINES 



347 



In general, when the requirements for exhaust steam are in excess of the 
steam consumption of a simple non-condensing engine a high-grade eco- 
nomical engine is without purpose. 

TABLE 58. 
ECONOMY OF AUTOMATIC SINGLE-CYLINDER NON-CONDENSING ENGINES. 





Pounds Steam per Indicated Horse Power Hour at 


Indicated Horse 






Power. 












Full Load. 


Three-quarter Load 


One-half Load. 


One-quarter Load. 


80 


29.42 


29.93 


31.66 


37.08 


100 


28.96 


29.40 


31.04 


36.00 


125 


28.47 


28.84 


30.42 


35.10 


150 


28.12 


28.46 


29.95 


34.38 


200 


27.51 


27.81 


29.25 


33.26 


300 


26.64 


26.75 


27.97 


31,68 



Above engine economies are based on dry saturated steam at 125 lbs. Engine 
horse powers are figured with steam cutting off at 25% of stroke. The economies 
given are for medium speed piston valve engines of the highest type only. 

202. High-speed Multi-valve Simple Engines. — The steam distribu- 
bution in a single-valve engine may give good economy for a very small 
range in load but be far from satisfactory for a wide range. This must 
necessarily be so since admission, cut-off, release, and compression are 
all functions of one valve, and any change in one results in a change of 
the others. To obviate the limitations of the single valve, many 
builders design engines with two or more valves. With a two-valve 
engine cut-off is independent of the other events, and with four valves 
all events are independently adjustable. In addition to the flexibility 
of the valve gear, the chief feature of the four-valve engines lies in the 
reduction of clearance volume which is made possible by placing the 
valves directly over the ports. The valves may be of the common 
slide-valve or rotary type. As a class, four-valve engines are more 
economical than those having a less number of valves. The advantages 
and disadvantages of the four-valve over the single-valve engines may 
be tabulated as below. 



Advantages. 
Better steam distribution. 
Better regulation. 
Reduced clearance volume. 
Less valve leakage. 
Better economy. 



Disadvantages. 

1. Increased number of parts. 

2. Increased first cost. 

3. Requires greater attention. 



The steam consumption of a high-speed four-valve non-condensing 
engine varies from 22 to 35 pounds of saturated steam per horse-power 



348 



STEAM POWER PLANT ENGINEERING 



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STEAM ENGINES 



349 



hour, with an average not far from 27 pounds. With superheated steam 
the steam consumption may run as low as 15 pounds per horse-power 
hour. 

Fig. 208 gives a comparison between a single-valve and a four-valve 
high-speed engine, and though the engines differ slightly in size, the 
conditions of operation were comparable and the marked gain in economy 
of the latter over the former is apparent. Both performances are ex- 
ceptional, and a 10 to 15 per cent greater steam consumption may be 
expected in average good practice. 

100 



90 



70 



ft 60 



50 



40 



20 















o 










Comparative Economy 












• of a 

( A ) Single Valve High Speed 
and a 






















( B) Four Valve High Speed 
Non Condensing Engine 












15 x 14 Keeves Simple ( A) 












16 x 16 Fleming Simple ( B ) 




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100 110 120 130 140 



10 20 30 40 50 60 70 80 90 
Per Cent of Rated Load 
Fig. 208. 

As a general rule single-valve simple engines do not exceed 500 horse 
power in size for stationary work, whereas 1000 horse power is not an 
uncommon size for the multi-valve type. 

203. Medium and Low-speed Multi-valve Simple Engines. — A com- 
parison of tests of high- and low-speed single-valve engines irrespective 
of design and construction shows the former as a class to be less eco- 
nomical than the latter. With four-valve engines there is no such dis- 
parity, and the high-speed type has shown just as good economy as the 
slow-speed class. For example, Engine No. 17, Table 59, with Corliss 
valves and a speed of 210 r.p.m., gives practically^ the same economy 
as Corliss engine No. 15 operating at 62 r.p.m. By far the greater num- 
ber of simple multi-valve slow-speed simple engines are of the Corliss 



350 



STEAM POWER PLANT ENGINEERING 



type. They range in size from 50 to 3000 horse power, with cylinders 
varying from 12 X 30 to 48 X 72. The smaller sizes with trip-valve 
gear run at 90 to 100 r.p.m., and the larger at 50 to 75 r.p.m. Without 
the trip gear, speeds of 150 r.p.m. are not uncommon, but at this speed 
they are usually classified as high-speed engines. 

Table 60 gives the steam consumption, condensing and non-condensing, 
of a number of four-valve slow-speed simple engines. For performances 
of uniflow engines, see paragraph 210. 

TABLE 60. 

ECONOMY OF CORLISS OR MEDIUM-SPEED FOUR-VALVE SIMPLE ENGINES. 

Following engine economies are based on dry saturated steam at 130 pounds and 
exhausting against atmospheric pressure. Engine horse powers are figured with 
steam cutting off in the cylinder at 25 per cent of stroke. 





Pounds Steam per Indicated-horse-power Hour at 


Indicated Horse 






Power. 












Full Load. 


Three-quarter Load 


One-half Load. 


One-quarter Load. 


200 


23.45 


23.05 


24.75 


35.00 


350 


23.03 


22.54 


24.07 


33.79 


500 


22.61 


22.06 


23.45 


32.73 


650 


22.24 


21.67 


22.92 


31.71 


800 


22.00 


21.40 


22.57 


30.91 


900 


21.90 


21.31 


22.44 


30.48 



204. Compound Engines. — Compound engines may be divided into 
three classes, tandem, cross compound, and duplex. In the tandem 
the two cylinders are end to end, in the cross compound side by side, 
and in the duplex one above the other. The tandem and duplex com- 
pounds have the advantage of (1) compactness for a given power, (2) 
less complication and fewer parts, and (3) low first cost. The crank 
effort is more variable than in the cross compound. In very large 
engines the low-pressure stage is generally divided between two cylinders 
of equivalent size to avoid an excessively large single cylinder and to 
distribute the crank effort. High-speed non-condensing compounds are 
ordinarily of the tandem type and are finding much favor in isolated 
station work, as in the power plants of tall office buildings where ground 
space is limited, though the duplex compound is sometimes used. The 
vertical or horizontal cross compound is generally installed in street- 
railway plants. 

Cylinder ratios for high-speed single-valve compound engines vary 
from about 1 to 2\ with 100 pounds pressure to about 1 to 3 with a 
pressure of 150 pounds, and for slow-speed condensing engines from 1 to 
3 with 125 pounds pressure to about 1 to 4 with a pressure of 175 pounds. 



STEAM ENGINES 



351 




Fig. 209. 3500 K.W. Vertical Cross-Compound Corliss Engine as Installed at the 
Power House of the Twin City Rapid Transit Co., Minneapolis, Minn. 



352 



STEAM POWER PLANT ENGINEERING 




Fig. 210. 7500 K.W. Vertical-horizontal Double-compound Engine as Installed at the 
59th Street Station of the Interborough Rapid Transit Co. (Manhattan Type.) 



STEAM ENGINES 353 

G. I. Rockwood recommends a ratio as high as 7 to 1, and a number of 
engines designed along this line have shown exceptional economy. A 
cross-compound Corliss engine at the Atlantic Mills, Providence, R.I., 
with cylinders 16 and 40 X 48 (ratio 6.128 to 1) gave the low steam 
consumption of 11.2 pounds of steam per i.h.p. hour, corresponding to a 
heat consumption of 222 B.t.u. per i.h.p. per minute. The 5500-horse- 
power engines of the New York Edison Company have a cylinder ratio 
of 6 to 1. The great majority of compound engines, however, have 
cylinder ratios of 4 to 1 or less. The 8000-horse-power engines of the 
Interborough Rapid Transit system have a ratio of 4 to 1, and the 4000- 
horse-power units of the Metropolitan Elevated Company, New York, a 
ratio of 3.5 to 1. 

The respective advantages and disadvantages of compounding may 
be tabulated as follows: 

Advantages. Disadvantages. 

1. Permits high range of expansion. 1. Increased first cost due to multi- 

2. Decreased cylinder condensation. plication of parts. 

3. Decreased clearance and leakage 2. Increased bulk. 

losses. 3. Increased complexity. 

4. Equalized crank effort. 4. Increased wear and tear. 

5. Increased economy in steam 5. Increased radiation loss. 

consumption. 

The ratio of expansion for a multi-expansion engine is usually taken 
to be the product of the ratio of the volume of large to small cylinder 
divided by the fraction of the stroke at cut-off in the high-pressure 
cylinder. For example, a compound engine with cylinders 24, 48 X 48 
cutting off at J in the high-pressure cylinder has a nominal ratio of 
expansion of 4 -f- J = 12. The number of expansions at rated load in 
compound condensing engines varies widely, ranging from 10 to 33, 
with an average not far from 16. 

The steam consumption shown by tests of a number of compound 
engines using saturated steam, condensing and non-condensing, is given 
in Table 66. For tests with superheated steam see Table 69. 

Fig. 211 shows the relative economy under comparable conditions 
of a high-speed simple and a high-speed compound engine, both running 
non-condensing and using saturated steam. The advantage of the 
compound at full load and overload is very marked, though its economy 
drops off rapidly at light loads and may be less than that of the simple 
engine. 

Fig. 212 shows the relative economy of two compound Corliss engines 
running condensing and non-condensing, both using saturated steam. 



354 



STEAM POWER PLANT ENGINEERING 



It should be borne in mind that the object of compounding is to 
permit the advantageous use of high pressures and large ratios of ex- 



25 



20 





\ 




































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\ 




































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Relative Economy 

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Non-Condensing High Speed 

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140 



160 



180 



pansion. Under proper conditions compounding may increase the 
economy at rated load about 20 per cent for non-condensing engines 
and 30 per cent for condensing engines. 



































1 


1 
1 


1 




































A 21,41 x 30 Compound. 
B 20,40 x 42 Compound 










































20 


















A Os 














































































18 




























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ensing 


-o- 






























































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14 






Bo 


s 




















































C 






































12 
















-— HSgdensing 


Z 








o 































































300 400 500 600 700 800 900 
Indicated Horse Power 
Fig. 212. 



1000 1100 1200 1300 1400 



An exceptional performance of a single-valve high-speed non-con- 
densing compound engine is that of engine No. 20, Table 66. With 
initial gauge pressure of 128 pounds the steam consumption is 22.3 
pounds per i.h.p. hour, corresponding to a heat consumption of 376 B.t.u. 
per i.h.p. per minute. 



STEAM ENGINES 



355 



One of the best performances of a multi-valve high-speed compound 
non-condensing engine is that of engine No. 14, Table 66. With initial 
pressure of 175 pounds gauge the steam consumption at full load is 
17.17 pounds per i.h.p. hour, corresponding to a heat consumption of 
291 B.t.u. per i.h.p. per minute. 

The 8000-horse-power vertical cross-compound Corliss engines of the 
Interborough Rapid Transit system (No. 6, Table 66) probably hold 




2000 2500 3000 
Gross K.W. Output 



3500 4000 4500 5000 



Fig. 213. 



Economy : Test of the 5500-Horse-power Three-cylinder Compound Engine and 
Generator at the Waterside Station of the New York Edison Co. 



the record for economy for compound engines without jackets and re- 
heaters, using saturated steam. With initial pressure of 175 pounds 
gauge and absolute back pressure of 2.2 pounds, the steam consumption 
is 11.96 pounds per i.h.p. hour, corresponding to a heat consumption 
of 220 B.t.u. per i.h.p. per minute. In estimating average practice it 
would be safe to add 10 per cent or 20 per cent to the steam consumptions 
given in Table 66. 

Fig. 213 illustrates the performance of the 5500-horse-power three- 
cylinder compound engine at the Waterside Station of the New York 
Edison Company. The best economy is 11.93 pounds of steam per 
i.h.p. hour, corresponding to a heat consumption of 221 B.t.u. per i.h.p. 
per minute. 



356 



STEAM POWER PLANT ENGINEERING 



TABLE 61. 

ECONOMY OF AUTOMATIC TANDEM COMPOUND ENGINES RUNNING NON- 
CONDENSING. 



Indicated Horse 


Pounds Steam per Indicated Horse Power at — 


Power. 


Full Load. 


Three-quarter Load. 


One-half Load. 


One-quarter Load. 


100 
150 
200 
250 
350 
450 


24.04 
22.94 
22.36 
21.98 
21.48 
21.27 


25.06 
23.82 
23.19 
22.75 

22.18 
21.92 


29.54 
28.00 
27.19 
26.65 
25.96 
25.61 


43.84 
41.40 
40.46 
39.13 
38.01 
37.45 



Above engine economies are based on dry saturated steam at 140 pounds. Cylin- 
der ratios are 4 to 1, and engine horse powers are figured with steam cutting off in the 
high-pressure cylinder at 25 per cent of stroke. The economies given are for medium- 
speed piston-valve engines of the highest type only. 

Above cylinder ratios are for engines normally operating condensing. These 
economies are given, however, so that engineers may know the steam consumption 
of this class of engine when it becomes necessary to operate same non-condensing, 
through lack of condensing water, or when it is desired to use the exhaust steam for 
heating purposes. 



TABLE 62. 

ECONOMY OF AUTOMATIC TANDEM COMPOUND CONDENSING ENGINES. 

Following engine economies are based on dry saturated steam at 140 pounds, 
and vacuum of 26 inches. Cylinder ratios are 4 to 1, and engine horse powers are 
figured with steam cutting off in the high-pressure cylinder at 25 per cent of stroke. 
The economies given are for medium-speed piston valve engines of the highest 
type only. 





Pounds Steam per Indicated Horse-power Hour at — 


Indicated Horse 




Power. 












Full Load. 


Three-quarter Load. 


One-half Load. 


One-quarter Load. 


150 


20.25 


21.51 


26.00 


37.83 


300 


19.10 


20.12 


24.10 


33.90 


400 


18.55 


19.45 


23.11 


32.07 


500 


18.15 


18.93 


22.44 


30.90 


600 


17.92 


18.67 


22.08 


30.20 


700 


17.83 


18.55 


21.91 


29.95 



STEAM ENGINES 



357 



TABLE 63. 

ECONOMY OF CORLISS OR MEDIUM-SPEED FOUR-VALVE TANDEM COMPOUND 
ENGINES RUNNING NON-CONDENSING. 



Indicated Horse 


Pounds Steam per Indicated Horse-power Hour at — 


Power. 


Full Load. 


Three-quarter Load. 


One-half Load. 


One-quarter Load. 


300 
450 
600 
750 
850 
950 


19.71 
19.20 

18.91 
18.74 
18.68 
18.66 


21.74 
21.10 
20.66 
20.41 
20.32 
20.30 


26.82 
25.90 
25.20 
24.73 
24.55 
24.48 


39.54 
37.80 
35.46 
35.31 
34.81 
34.47 



Above engine economies are based on dry saturated steam at 150 pounds pressure 
and exhausting at atmospheric pressure. Cylinder ratios are 4 to 1, and engine 
horse powers are figured with steam cutting off in the high-pressure cylinder at 25 per 
cent of stroke. 

Above cylinder ratios are for engines normally operating condensing. These 
economies are given, however, so that engineers may know the steam consumption 
of this class of engine when it becomes necessary to operate same non-condensing, 
through lack of condensing water, or when it is desired to use the exhaust steam for 
heating purposes. 



TABLE 64. 

ECONOMY OF CORLISS OR MEDIUM-SPEED FOUR-VALVE TANDEM COMPOUND 
CONDENSING ENGINES. 



Indicated Horse 


Pounds Steam per Indicated Horse-power Hour at — 


Power. 


Full Load. 


Three-quarter Load. 


One-half Load. 


One-quarter Load. 


300 

500 

700 

900 

1100 

1300 

1500 


15.42 
14.74 
14.29 
13.97 
13.73 
13.56 
13 .49 


15.30 
14.60 
14.10 
13.76 
13.51 
13.33 
13.23 


17.26 
16.45 
15.78 
15.34 
15.02 
14.83 
14.71 


24.00 
22.47 
21.30 
20.48 
19.94 
19.66 
19.52 



Above engine economies are based on dry saturated steam at 160 pounds and 
26 inches vacuum. Cylinder ratios are 4 to 1 and engine horse powers are figured 
with steam cutting off in the high-pressure cylinder at 25 per cent of stroke. 



358 



STEAM POWER PLANT ENGINEERING 




J 
s\ 

7 ( 

^-------3 6 

\ '2d 

: - 

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a 

i i a 

i S 

i^S^ a 



STEAM ENGINES 



359 



205. Triple and Quadruple Expansion Engines. — Triple and quadruple 
expansion engines are still in use where the load is practically constant, 
as in marine and pumping-station practice, but have been abandoned 
in street-railway work and in plants where the load fluctuates widely 

TABLE 65. 
ECONOMY OF MODERN VERTICAL TRIPLE-EXPANSION PUMPING ENGINES. 

(Official Trials.) 





Type. 


Location. 


Rated 
Capacity, 
Millions 
of U.S. 
Gallons. 


Initial 

Gauge 

Pressure. 


Duty. 


Dry 


Date of 
Test. 


Per Thou- 
sand Lbs. 
of Dry 
Steam. 


Per 

One 

Million 

B.t.u. 


Steam 

per 
I.h.p. 
Hour. 


5- 2-09 
3-10-10 


Holly 
Holly 
Holly 

Holly 
Holly 

Allis 

Allis 
Allis 
Allis 


Louisville, Ky 

Frankfort, Pa 

Albany, N. Y 

Brockton, Mass 

Cleveland, Ohio 

Boston, Mass 

St. Louis, Mo 

St. Louis, Mo 

Milwaukee, Wis 


24 
20 
12 

6 

2.5 
30 

20 
15 
12 


155.1 

180.2 
153.0 

150.0 
149.6 
185.5 

140.6 
126.2 
124.6 


*195.0 
184.4 
182.1 

170.0 
164.6 

178.5 

181.3 
179.4 
175.4 


164.5 


*9.64 


4-29-10 






10-14-09 






12- 5-07 
5- 2-00 

2- 4-06 
2-26-00 
1-15-10 


148.8 
163.9 

158.8 
158.1 
151.0 


11.51 
10.33 

10.66 
10.67 
10.82 









109 degrees F. superheat at throttle. 






Date of 
Test. 


Type. 


R.P.M. 


Water Actually 
Pumped, Mil- 
lions of U.S. 
Gallons 24 Hrs. 


Net Head 

Pumped 

Against, Lbs. 

per Sq. In. 


Indicated 
Horse 
Power. 


Developed 
Horse 
Power. 


Thermal 

Efficiency 

Per Cent. 

I.h.p. 


5- 2-09 
3-10-10 
4-29-10 


Holly 
Holly 
Holly 


24.0 
20.1 
22.3 


24.111 
21.219 
12.193 


90.0 

95.7 

139.5 


925.7 


879.4 
817.0 
726.0 


22.54 


10-14-09 

12- 5-07 

5- 2-00 


Holly 
Holly 
Allis 


40.1 
62.3 
17.7 


6.316 

2.142 

30.314 


130.6 

180.7 

61.0 


158^7 
801.5 


334.0 
151.9 

747.8 


19' 13 
21.63 


2- 4-06 
2-26-00 
1-15-10 


Allis 
Allis 
Allis 


16.5 
16.4 
20.4 


20.070 
15.121 
12.430 


104.0 
127.0 
121.0 


859.2 
801.6 
673.0 


839.6 
726.3 
618.0 


20.92 
21.00 
20.25 



in favor of the two- or three-cylinder compound. The best economy on 
a heat-unit basis ever recorded for an engine using saturated steam was 
that of the Nordburg quadruple pumping engine at Wildwood, Pa., 
which gave a consumption of 12.26 pounds per i.h.p. hour and a heat 
consumption of 186 B.t.u. per i.h.p. per minute reckoned above the 
feed- water temperature.* The Allis triple-expansion pumping engine 

* Replaced in 1905 by a Riedler pumping engine on account of high maintenance 
cost. 












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STEAM ENGINES 361 

at Chestnut Hill holds the record for saturated -steam consumption, 
10 pounds per i.h.p. hour, and its exceptional performance of one de- 
veloped horse power per 1.09 pounds of coal has, perhaps, never been 
excelled. 

Triple Expansion Engines. — Cylinder Proportions for Triple Expansion Engines: 
Trans.' A.S.M.E., 21-1002, 10-576. Economy of Triple Expansion Engines: Trans. 
A.S.M.E., 8-496. 

206. Effects of Condensing. — The effect of the condenser upon the 
power and economy of engines is indicated in Table 67. The curves 
in Figs. 215 and 216 were plotted from tests made by Professor R. L. 
Weighton on a 7, 10J, 15§ X 18 triple-expansion engine at Durham 
College of Science, Newcastle-on-Tyne. The straight line shows how 
the mean effective pressure would vary with the degree of vacuum if 
the power increased directly with the reduction in back pressure. The 
curved line shows the actual m.e.p., which increases almost along the 
theoretical line up to a 10-inch vacuum, from which point on the in- 
crease is less marked. At 26 inches the actual m.e.p. reaches an 
apparent maximum. These figures are not applicable to all engines but 
give a good idea of the limitation of the vacuum with the average type 
of reciprocating engine with restricted exhaust port openings. With 
specially designed ports and passages of large cross-sectional area the 
piston engine shows increase in steam economy up to the highest vacuum 
carried in the condenser. (See Power, Jan. 16, 1912, p. 72.) 

The gain in steam consumption due to the condenser does not indicate 
a corresponding gain in heat consumption. For example, Engine No. 2, 
Table 67, shows an apparent gain in steam consumption, due to con- 
densing, of 12.5 per cent, the temperature of the feed water returned 
to the boiler being 120 degrees F. With a suitable heater the exhaust of 
the non-condensing engine would be capable of heating the feed water 
to 210 degrees F. The non-condensing engine should therefore be 
credited with 210 — 120 or 90 heat units per pound of steam used, or, 
in round numbers, 9 per cent. The difference between 12.5 per cent 
and 9 per cent, or 3.5 per cent, represents the net gain in favor of con- 
densing, provided the power necessary to create the vacuum is ignored. 
Actually, the steam consumption of the condenser pumps might be 
equal to or greater than 3.5 per cent of the steam generated and the 
net gain becomes zero or even negative. Referring to Fig. 217, plotted 
from tests of the 7, 10J, 15J X 18 triple-expansion engine mentioned 
above, the solid lines show the feed-water consumption per i.h.p. hour 
and the broken line the heat units consumed per brake horse power per 
minute measured above the hot-well temperature. The engine effi- 
ciency, based upon the water consumption, increases as the vacuum 



362 



STEAM POWER PLANT ENGINEERING 



£ 36 
"£. 

S « 
« 34 

I 33 
« 32 
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^ 30 









































































































































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Inc 


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pansi 


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uum 



































































































8 10 12 14 16 18 20 
Vacuum in Inches of Mercury 

Fig. 215. 



22 24 




10 12 14 16 18 20 22- 
Vacuum in Inches of Mercury 



. 26 28 30 



Fig. 216. 



STEAM ENGINES 



363 



increases, reaching a maximum between 26 and 28 inches, whereas the 
heat-unit curve gives the maximum between 20 and 21 inches. Be- 
tween 22 and 28 inches the heat-unit curve shows a rapid falling off 
in economy. Tests of the 5500-horse-power engine at the New York 
Edison Company's Waterside Station showed that increasing the vacuum 
from'25.3 to 27.3 inches decreased the water rate only 0.06 pound per 

TABLE 67. 

EXAMPLES OF THE EFFECT OF CONDENSING ON THE ECONOMY OF 
RECIPROCATING ENGINES. 





Non-Condensing. 


Condensing. 


Increase Due 
to Condensing. 


8 J 

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a 


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54.7 


19.2 


149 


1.6 


83.4 


14.8 


52.5 


25 


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148 


540 


19.3 


147 


4 




16.9 


* 


12.5 


3 


126 


83 


23.8 


130 


7.4 


116 


19.1 


39.8 


19.7 


4 


67.6 


209 


28.9 


67 


4.5 


213 


22 


1.9 


23.5 


5 


103.8 


177.5 


22.1 


103.8 


1.2 


155 


16.5 


* 


25.1 


6 


114 


160 


31 


114 




168 


27 


2 


12.9 


7 


96 


120 


23.9 


96 


4 


145 


19.4 


20.8 


18.8 


8 


118 


267 


23.24 


119 


4.2 


276.9 


16 


3.7 


31 


9 


75.9 


310 


25.6 


79 


6.4 


336 


20.5 


8.7 


19.9 


10 


62.5 


451 


30.1 


63.6 


7.8 


444 


23 


* 


23.6 


11 


186.7 


40.4 


18.7 


184.6 


1.6 


29.8 


12.7 


* 


32 



* Cut-off changed for best economy. 

1. 7, 10J, 15£ x 18 triple ; Eng. News, Aug. 21, 1902, p. 127. 

2. 17, 27 x 24 Westinghouse marine, non-condensing ; Power, August, 1903. 

3. 1, 18 x 10 Buffalo tandem compound ; Elec. World, Sept. 10, 1904, p. 404. 

4. 18 x 30 four-valve (slide) ; Engine Tests, Barrus, p. 88. 

5. 21, 65 x 43.31 Corliss ; Peabody's Thermodynamics, p. 382. 

6. 12 x 12 Reeves simple ; Elec. World, Oct. 1, 1904, p. 587. 

7. 18 x 48 simple Corliss ; Peabody's Thermodynamics, p. 354. 

8. 14, 28 x 24 two- valve (slide) ; Engine Tests, Barrus, p. 175. 

9. 17 x 24 two-valve; Engine Tests, Barrus, p. 70. 

0. 28 x 36 Corliss ; Engine Tests, Barrus, p. 97. 

1. Willans triple expansion central valve engine ; Peabody, Thermodynamics, p. 406. 



i.h.p. hour. (Power, July, 1904, p. 424.) The results are illustrated in 
Fig. 217. In most cases, and particularly with large compound engines, 
the net gain due to condensing is considerable, but the feed-water 
temperatures and power consumed by the auxiliaries should be taken 
into account.* Fig. 206 shows the effect of vacuum on the steam con- 
sumption of a small high-speed simple engine, and Fig. 212 of a cross- 
compound Corliss. (See also paragraph 236.) 

* See Power, Feb. 23, 1909, p. 381. 



364 



STEAM POWER PLANT ENGINEERING 



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Fig. 217. Performance of 5500-H.P. Engine at Waterside Station of New York Edison 

Company. 

207. Throttling vs. Automatic Cut-Off. — The action of the gov- 
ernor in the throttling engine is shown by the superposed indicator 
cards (Fig. 218) taken between zero or friction load and maximum 
load. The effect of throttling is to reduce the pressure during admis- 
sion, but does not change the point of cut-off or other events of the 




Fig. 218. Typical Indicator Cards. High-speed Throttling Engine. 

stroke. The steam may be partially dried or even superheated by 
throttling, thus tending to reduce cylinder condensation. Initially 
dry saturated steam at a pressure of 125 pounds gauge would be super- 
heated about 12 degrees in expanding through a throttle to 90 pounds, 
or if it contained initially 2 per cent moisture would be perfectly dried 
in expanding to 40 pounds. (See Table 68.) Friction through the 
valve also tends to dry the steam. Thus with very light loads the 



STEAM ENGINES 



365 



superheat may be decidedly appreciable. The possible gain due to 
decreased cylinder condensation is to some extent offset by incomplete 
expansion. The best efficiency for a given load is realized by a proper 
compromise between cut-off and initial pressure. Experiments made 
by Professor Denton (Trans. A.S.M.E., 2-150) on a 17 X 30 non- 
condensing double-valve engine showed the most economical results 
with \ cut-off for 90 pounds pressure, \ cut-off for 60 pounds, and T 4 o 5 o 
for 30 pounds. The average throttling engine does not give close 
regulation, the governor usually lacking sensitiveness. Tests show the 
economy to be better than that of the automatic engine on light loads, 
and the crank effort more uniform. 

TABLE 68. 

SHOWING THE INITIAL PER CENT OF MOISTURE THAT WILL BE EVAPORATED 

IN THROTTLING FROM A HIGHER TO A LOWER PRESSURE. 

Based on Marks' and Davis' Steam Tables. 



Final Pressures. 



Final Pressures. 



80 




75 


0.14 


70 


0.28 


65 


0.43 


60 


0.59 


55 


0.77 


50 


0.97 


45 


1.19 


40 


1.44 


35 


1.72 


30 


2.05 


25 


2.44 


20 


2.90 


15 


3.51 



Initial Pressure, Absolute. 



SO 



85 



0.13 
0.26 
0.40 
0.55 
0.71 
0.89 
1.09 
1.32 
1.56 
1.85 
2.18 
2.56 
3.04 
3.65 



90 



0.24 
0.37 
0.52 
0.66 
0.83 
1.01 
1.21 
1.44 
1.68 
1.97 
2.30 
2.69 
3.16 
3.78 



0.36 
0.49 
0.64 
0.78 
0.95 



100 



0.45 
0.59 
0.74 
0.88 
1.06 



23 
,43 
66 
91 
20 
.53 



2.92 
3.40 
4.01 



105 



0.55 
0.70 
0.84 
0.99 
1.16 
1.34 
1.54 
1.76 
2.02 
2.31 
2.64 
3.03 
3.51 
4.13 



no 



2.12 
2.41 
2.74 
3.13 
3.61 
4.23 



115 



0.74 

0.88 

1.03 

18 

34 

53 

74 

96 

21 

2.51 

2.84 

3.23 

3.71 

4.33 



80 
75 
70 
65 
60 
55 
50 
45 
40 
35 
30 
25 
20 
15 



Initial Pressure, Absolute. 



125 



0.91 

1.05 

1.19 

1.34 

1.52 

1.70 

1.90 

2.13 

2.39 

68 

.01 

41 

88 

51 



130 



1.99 



2.21 
2.47 
2.77 
3.10 
3.49 
3.97 
4.60 



135 



1.08 



1, 

1, 

1, 

1. 

1.86 

2.08 



21 
36 
51 
.68 



4.06 
4.70 



140 145 



1.15 



1.28 
1.43 
1.59 
1.76 
1.94 
2.15 
2.38 
2.63 
2.93 
3.26 
3.66 
4.15 
4.78 



1.22 
1.36 
1.50 
1.66 
1.83 
2.02 
2.22 
2.45 



150 



1.29 
1.43 
1.58 
1.73 
1.90 
2.09 
2.30 
2.52 
2.78 
3.08 
3.41 
3.81 
4.30 
4.94 



155 



1.35 



49 
64 
79 
96 
15 
36 
59 
84 



3.14 
3.48 
3.88 
4.37 
5.01 



160 



1.41 



55 

70 

85 

03 

21 

2.42 

2.65 

2.91 

3.21 

3.55 

3.96 

4.45 

5.09 



120 



0.83 
0.97 
1.12 
1.26 



44 
62 
82 
05 
30 
60 
93 
32 



3.80 
4.43 



165 



1.48 
1.62 
1.77 
1.93 
2.10 
2.29 
2.50 
2.73 
2.99 
3.29 
3.63 
4.04 
4.53 
5.17 



366 



STEAM POWER PLANT ENGINEERING 



j The indicator cards shown in Fig. 219 were taken from a single- 
valve high-speed automatic engine operating between friction load and 
maximum load. The mean effective pressure is adjusted to suit the 
load by the automatic variation in the cut-off, the initial pressure 
remaining the same. Since the cut-off is controlled by the action of 
the governor on the single valve, all other events of the stroke are like- 
wise changed. With a four-valve engine the variation in cut-off does 
not affect the other events. 




Fig. 219. Typical Indicator Cards. High-speed Automatic Engine. 

The chief advantage of the automatic over the throttling engine lies 
in its sensitive regulation, and while, in general, it gives a lower steam 
consumption than the throttling engine, this is probably in most cases 
due to superior construction and not to the method of governing. 

The following performances of a Belliss 250-horse-power high-speed 
condensing engine fitted with both automatic and throttling governing 
devices give results decidedly in favor of the throttling engine. (Pro. 
Inst, of Mech. Engrs., 1897, p. 331.) 



Percentage of load .... 
Electrical horse power, 
Steam per I.H.P. hour 



Automatic Cut-Off. 




Throttling. 


100 


62.5 


33 


25 


100 


62.5 


33 


213 


132 


77.8 


53 


213 


132 


77.8 


22.5 


22.9 


28.5 


34.3 


21 


21.7 


25.6 



25 
53 
28.4 



Some of the comparative advantages and disadvantages of the auto- 
matic and throttling engines are as follows: 



Automatic. 

Advantages. 
Sensitiveness of regulation. 1. 

Increased ratio of expansion. 2. 

Low terminal pressures. 3. 

4. 
Disadvantages. 
Increased cylinder condensation. 1. 

Greater variation in crank effort. 2. 

Complicated valve gear. 3. 

Low economy at very early loads. 



Throttling. 

Low first cost. 
Crank effort more uniform. 
Reduced cylinder condensation. 
Simplicity of regulating device. 

Low ratio of expansion. 
High terminal pressure. 
Low initial pressure at early loads. 



STEAM ENGINES 



367 




Mean Pressure on L.P.Piston, Lb.per Sq.In. 

Fig. 220. Throttling vs. Automatic Gut-off. 



Fig. 220 shows the relative steam consumptions of an engine under 
the same conditions of load when controlled by variable expansion and 
by throttling. Suppose this 
engine to be altered in capacity 
so that the m.e.p. referred to 
the low-pressure piston is about 
32, then the steam consumption 
with the throttling governor will 
be as shown by straight line A. 
This shows that between 32 and 
12 pounds m.e.p. very little is 
gained by a variable expan- 
sion, and below that load the 
throttled governor is the more 
economical. (Power, Feb. 21, 
1911, p. 301.) 

208. Influence of Superheat. — (See also paragraph 117.) Table 69 
gives test results for several different types of engine employing super- 
heated steam. These figures may be compared with the performances 
of engines using saturated steam as given in Tables 59 and 66. A 
decided gain in economy is shown in favor of superheat for single-cylinder 
engines. With compound engines the advantage is not so apparent, 
while triple-expansion engines show the least gain. Tables 69 to 72 
show the effect of superheating on simple, compound and triple-expan- 
sion engines and represent average current practice. (Proc. A.S.M.E., 
September, 1907.) Better results than these have been obtained with 
such engines as the Lentz compound, but for ordinary superheated steam 
practice the results in the tables may be used with confidence. Some 
idea of the wonderful fuel economy effected in Europe with the use of 
highly superheated steam in connection with the so-called locomobile 
is gained from the results shown in Table 73. This type of engine has 
not yet been introduced to any extent in this country but it is only a 
matter of time when the cost of coal will advance to such a point as to 
preclude all but the more economical types of prime movers. 

As far as steam consumption is concerned, all engines show greater 
economy with superheated than with saturated steam, but the thermal 
gain is not so marked, and when the economy is measured in dollars 
and cents per developed horse power, taking all things into consideration 
the gain is still further reduced and in some cases completely neutralized. 
First cost, maintenance, and disposition of the exhaust must all be 
considered in determining the ultimate commercial gain due to the 
use of superheated steam. 



368 



STEAM POWER PLANT ENGINEERING 



Fig. 221 gives the results of a series of tests made on a number of 
Belliss & Morcom engines using superheated steam. (Pro. Inst, of 
Mech. Engrs., March, 1905, p. 302.) The engines were from 200 to 
1500 kilowatts capacity and were tested at full load. It is noticeable 
that the curves all converge to a single point and will meet at about 
400 degrees F. The results show that if sufficient superheat is put into 



25 



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300 



200 

Superheat, Deg. F. 
Fig. 221. Effect of Superheat on Steam Consumption. 



400 



the steam all engines of whatever size are equally economical. Fig. 222 
shows the relationship between degree of superheat and the heat con- 
sumption at various loads for a 300-horse-power Belliss & Morcom high- 
speed triple-expansion engine. (Pro. Inst. Mech. Engrs., March, 1905, 
p. 303.) It will be noted that the variation in heat consumption at 
different percentages of load becomes less marked as the degree of 
superheat increases. With superheat of 350 degrees F. the heat con- 
sumption from i load to full load is practically constant. 



STEAM ENGINES 



369 



These curves though strictly applicable to the specific cases cited 
are more or less general and represent the influence of superheat on all 
types of piston engines. 



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100 200 

Degrees of Superheat, Fahrenheit 
Fig. 223. 



370 



STEAM POWER PLANT ENGINEERING 




Fig. 224. 3000 H.P. Sulzer Engine Designed for Highly Superheated Steam. 



STEAM ENGINES 



371 




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376 



STEAM POWER PLANT ENGINEERING 



209. The Locomobile. — Fig. 226 shows a section through the engine 
cylinders and the boiler setting of a Wolf tandem-compound locomobile, 
illustrating a type of steam plant which is finding much favor in Europe. 
The entire equipment is self-contained and requires very little floor 
space. The engine is set upon the boilers with the cylinders projecting 
into the smoke box so as to minimize piping and radiation losses. Steam 
is generated at a pressure of 175-225 pounds absolute and is superheated 
by the furnace gases to 800-850 degrees F. before it is admitted to the 
high-pressure cylinder. Exhaust steam from the high-pressure cylinder 




Wolf Tandem-compound Locomobile. 



is reheated by an auxiliary superheater (adjoining the main super- 
heater) to a temperature of 450-550 degrees F. before it enters the low- 
pressure cylinder. The feed water is heated by an economizer placed 
in the breeching. All auxiliaries are driven by the main engine. 

Locomobiles are made in various sizes ranging from 25 to 1000 horse 
power and are designed for condensing or non-condensing operation. 
Coal consumptions as low as 0.8 pound per developed horse-power hour 
have been realized and one pound per horse-power hour is common 
practice. On account of the extremely high degree of superheat em- 
ployed the load curve is very flat and there is little difference in economy 
between the small and large units. 

Table 73 gives the results of tests of a 100-horse-power Wolf tandem- 
compound locomobile and illustrates the remarkable economy which 
is being effected in Europe with this type of plant. 

The lowest steam consumption recorded to date, 6.95 pounds per 
indicated horse-power hour, is credited to an engine of this class. (See 
Zeit. d. Ver. Deut. Ingr., Mar. 18, 1911, p. 415.) 



STEAM ENGINES 



377 



TABLE 73. 

REMARKABLE ENGINE PERFORMANCE.' 

200, 400 X 400 mm. Locomobile. 









(7.8, 


15.7 X 15.7 in.) 












Steam Temperatures, Degrees F. 






Initial Pres- 
sure, Lbs. 
per Sq. In. 
Abs. 


Condenser 






Num-' 












ber of 


Pressure, 


Entering 


Leaving 


Entering 




R.P.M. 


Test. 


Lbs. Abs. 


High- 


High- 


Low- 


Final Feed 










pressure 


pressure 


pressure 


Water. 










Cylinder. 


Cylinder. 


Cylinder. 








220 


Condensing with Intermediate Superheating. 




1 


1.47 




Saturated. 




242 


236 


2 


227 


1.17 


712 


377 


462 


212 


241 


3 


220 


1.17 


718 


367 


460 


206 


242 


4 


221 


1.17 


806 


426 


530 


221 


246 


5 


220 


1.17 


842 


469 


538 




'243 


6 


220 


1.17 


872 


520 




241 


243 






Non-condensing without Intermediate Superheating. 




7 


220 




832 




462 


289 


237 


8 


220 




856 




505 


284 


238 


9 


221 




878 




527 


284 


242 


10 


220 




869 




572 


257 


241 


11 


220 




817 




525 


248 


241 


12 


221 




878 




568 


259 


241 



* Compiled from Zeit. des Ver. deut. Ingr., June, 1911. 











Steam Consumption, 






No. of 
Test. 


I.h.p. 


D.h.p. 


Mechanical 
Efficiency, 


Pounds 


Coal Burned, 
Lbs. per D.h.p. 


Heat Con- 

sumption.B.t.u. 

per I.h.p. per 

Minute. * 










Per Cent. 


Per I.h.p. 
Hr. 


Per D.h.p. 
Hr. 


Hr. 




112.5 


Condensing with Intermediate Superheating. 




1 


103.2 


91.6 


13.98 


14.19 


1.59 


260 


2 


138.4 


132.8 


96.0 


8.51 


8.87 


1.00 


198 


3 


140.3 


131.4 


93.5 


8.33 


8.90 


1.00 


195 


4 


140.4 


133.4 


95.0 


7.68 


8.06 


0.96 


186 


5 


138.8 


132.5 


95.5 


7.24 


7.56 


0.87 


175 


6 


141.8 


134.0 


94.5 


7.15 


7.56 


0.86 


175 




61.5 


Non-condensing without Intermediate Superheating. 




7 


49.3 


78.0 


11.22 


14.43 


1.65 


262 


8 


83.8 


74.0 


88.0 


10.60 


11.84 


1.17 


249 


9 


111.0 


98.5 


88.0 


9.95 


11.38 


1.12 


237 


10 


129.9 


120.8 


93.0 


10.00 


10.88 


1.07 


238 


11 


140.4 


132.2 


94.0 


10.68 


11.34 


1.12 


248 


12 


142.1 


132.4 


93.0 


9.93 


10.66 


1.05 


235 



Above ideal feed-water temperature corresponding to exhaust pressure. 



378 



STEAM POWER PLANT ENGINEERING 



210. Uniflow or Straight-flow Engine. — Fig. 227 shows a longitu- 
dinal section through a single-cylinder engine designed by Professor J. 
Stumpf, of Charlottenberg College, Germany, which is finding much 
favor with the European engineers, over a half-million horse power 
being in service at this writing. A 300-horse-power engine of this design 
is credited with a steam consumption of 8.5 pounds per i.h.p. hour; 
initial pressure 130 pounds absolute; superheat 261 degrees F., a per- 
formance equalled only by the best compound and triple-expansion 
engines. The engine is similar in principle to the two-cycle gas engine 
in which the elongated piston acts as an exhaust valve, opening and 
closing a series of slots in the middle of the cylinder shell. The live 
steam enters the engine through the cylinder head, which it heats, and 
is admitted into the cylinder through a double-seated poppet valve, 
and, after expansion, is exhausted through the slots in the middle of 




Fig. 227. Longitudinal Section through Stumpf Straight-flow Engine. 

the cylinder. A high degree of expansion is possible without excessive 
cylinder condensation, since the exhaust steam does not come into 
contact with the live-steam-heated surfaces of the cylinder as in the 
ordinary type of piston engine. 

On account of the high degree of expansion used in this type of engine 
the cylinders are necessarily large per unit output. The high compres- 
sion necessitates the use of heavy cylinders and flywheel and the inertia 
of the reciprocating parts requires a massive foundation. The cost per 
unit of power is therefore higher than that of the ordinary single-cylinder 
high-speed engine. The cost is, however, less than that of slow-speed 
compound and triple-expansion engines with which it successfully com- 
petes. Table 74 gives a comparison of the results of triple-expansion 
engines and a number of uniflow engines. 

The uniflow engine has not been generally adopted in this country 
notwithstanding its excellent showing in Europe. 

Fig. 228 shows a section through a uniflow engine as manufactured by 
the Nordberg Manufacturing Company, Milwaukee, Wis. The curves 
in Fig. 229 are based upon the tests of a 20 X 30, 200-horse-power 



STEAM ENGINES 



379 



Nordberg uniflow engine using very wet steam. The dotted lines show 
assumed results within the capacity of the machine. 

E 




W//////////////////////////y ^ 



wmm*^ 



&a 




I m/////////m^^^^ 



m 



Fig. 228. Section through Cylinder of Nordberg Uniflow Engine. 

TABLE 74. 

COMPARISON OF OPERATING PERFORMANCES OF TRIPLE-EXPANSION STEAM 

ENGINES AND STRAIGHT-FLOW STEAM ENGINES. 

(Power, Jan. 31, 1911.) 



Manufacturers. 





Diameter of Cyl- 






>" 


Steam. 


© 

1 


inders in Inches. 




c 


§ 






§ 






"dPt, 
















2© 




© 

s-g 


is 


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o a 


© 

ft 

s 


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2 


© © 




« £H 


a 


w 


tf 


a 


di 


H 



Steam Con- 
sumption 
per 



© fa 



Sulzer, Switzerland. . 
Gorlitz , Germany . . . 
Nurnberg, Germany. 



Best Performance of Stationary Triple-expansion Engines 


. 


6000 
6000 
6000 


40 
40 
41 


60 
60 
60 


2X73 
2X73 
2X73 


67 
67 
67 


83 
83 
83 


228 


170 
170 
170 


572 
572 
572 


8.5 
8.5 
8.5 


Straight-flow Steam Engines. 



Sulzer, Switzerland. . . . 

Same engine 

Gebr. Stock, Holland. 
Burmeister, Denmark. 



300 
300 
80 



23.5 
23.5 
12.6 
17.8 



31.5 
31.5 



23.5 



155 
155 
200 



130 
130 
149 
138 



617 




662 



8.5 
10.6 
9.7 
9.5 



13.7 
13.7 
13.7 



13.7 



211. Binary-vapor Engines. — A consideration of the Carnot or 
Rankine cycles shows that theoretically the efficiency of the steam 
engine may be increased by raising the temperature of the steam sup- 
plied or by lowering the temperature of the exhaust, that is to say, 
by increasing the range. Superheated steam development has prac- 
tically determined the upper limit, and economical practice indicates a 
vacuum of about 26 inches, corresponding to 126 degrees F., as the 
average lower limit for most efficient results from a commercial stand- 
point. 



380 



STEAM POWER PLANT ENGINEERING 



In the binary-vapor engine the working range has been considerably 
increased by substituting a highly volatile liquid, as sulphur dioxide, 
for the water which is ordinarily used as the cooling medium in the 
surface condenser. 

The S0 2 in condensing the exhaust steam is itself vaporized and the 
vapor, under a pressure of about 175 pounds per square inch, used 
expansively in a secondary reciprocating engine. The exhausted S0 2 
is discharged into a surface condenser in which it is liquefied by cooling 
water much the same as in refrigerating practice and used over and 
over again. Referring to Fig. 230, which illustrates diagrammatically 
a binary-vapor engine at the Royal Technical High School, Berlin: 



36 

31 
































































































«* 


32 

30 

28 
26 
24 
22 
20 
18 
16 
14 










































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Initial Pressure 150 fr Gauge 

150 R.P.M. 

Steam very wet 


— 










































































































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12 
10 











































































































50 



100 150 200 

Indicated Horse Power 



250 



300 



Fig. 229. Performance of Nordberg Uniflow Engine. 



A, B, and C are the three steam cylinders of an ordinary triple-expan- 
sion engine and D the S0 2 cylinder. All four cylinders drive a common 
crank shaft E. F is a high-pressure surface condenser which acts as 
a vaporizer for the S0 2 and a condenser for the steam. G is a surface 
condenser which serves to condense the S0 2 vapor. H is a liquid S0 2 
tank. The operation is as follows: Highly superheated steam enters 
the high-pressure steam cylinder at I and leaves the low-pressure 
cylinder at J, just as in any steam engine. The exhaust steam enters 
chamber F and is condensed by the liquid S0 2 passing through the 
coils. The condensed steam and entrained air are removed from the 
chamber by a suitable air pump. The steam in condensing gives up 
its latent heat to the liquid S0 2 and causes it to vaporize. The S0 2 
vapor passes from the coils in chamber F to the S0 2 engine D and 



STEAM ENGINES 



381 



performs work. The exhausted S0 2 vapor flows from cylinder D to 
chamber G, and is condensed by cooling water flowing through a series 
of tubes. The liquid S0 2 is collected in liquid tank H and thence is 
pumped into the coils in vaporizer F. The approximate temperatures 
and pressures at different points of the cycle are indicated on the 
diagram. 



-To Air_Pump 



37.6° 



S0 2 Vaporizer and Steam 
Condenser 
187 ^Absolute 148° F 



49.8 F 

Circulating 

Water 

Inlet 






Circulating 
Water 
Outlet 



Absolute, 
70°F 



Liquid 
SO„Tank 



Condenser 



4,2#-ABsT i 

o 
155 P | 
J I 



S0 2 Vapor 



L.P. 

Cylinder 



42.4 I.H.P. 



l.P. 

Cylinder 

1 



43.2 I.H.P. 



__590°F 



173frAbsolute 




56.8 I.H.P. 



SO 2 

Cylinder 



143.5 R.P.M. 



72°F 



S0 Exhaust 



Fig. 230. Diagram of Binary-vapor Engine. 



A number of experiments made by Professor E. Josse in the labora- 
tory of the Royal Technical High School of Berlin on an experimental 
plant of about 200 horse power gave some remarkable results. A few 
of the tests made with highly superheated steam gave the following 
average figures: 

I.h.p. (steam end) 146.4 

Steam consumption per i.h.p. hour 12.8 

I.h.p. (S0 2 end) 52.7 

Percentage of power of S0 2 engine 35.9 

Steam consumption per i.h.p. hour of combined engine 9.43 

When operating under the most satisfactory conditions a perform- 
ance of 8.36 pounds of steam per i.h.p. hour was recorded, correspond- 
ing to a heat consumption of 158.3 B.t.u. per minute, which is the 
best recorded performance to date (1912) in the history of steam- 
engine economy. 

S0 2 does not attack the metal surface of the engine unless combined 
with water, in which case sulphurous acid is formed. There is, how- 
ever, no danger from this cause, since the SO2 being under greater 
pressure effectually prevents leakage of water into the S0 2 system. 

The S0 2 cylinder requires no other lubrication than the S0 2 itself, 
which is of a greasy nature. 



382 



STEAM POWER PLANT ENGINEERING 



Properties of S0 2 : Trans. A.S.M.E., 25-181. Binary-Vapor Engines: Jour. 
Frank. Inst., June, 1903; Elec. World and Engr., Aug. 10, 1901; U. S. Cons. Re- 
ports, No. 1139, Sept. 14, 1901; Engr. U. S., Aug. 1, 1903; Sib. Jour, of Eng., March 
1902. 

212. Rotary Engines. — The rotary engine differs from the recip- 
rocating engine in that the piston, or equivalent, rotates about the 
cylinder axis. Its operation is entirely different from that of the 
steam turbine; in the rotary engine the static pressure of the steam 
actuates the piston and in the turbine the momentum of the steam is 
imparted to the rotating element. 




Fig. 231 Herrick Rotary Engine. 

Over 2200 patents have been issued to date on rotary engines but 
not a single machine has yet been able to compete with the reciprocat- 
ing engine as regards steam economy. The advantages of the rotary 
engine are many and for this reason innumerable inventors have been 
exerting their skill in the development of this type of prime mover, 
but unfortunately the impracticability of satisfactorily packing the 
rubbing surfaces has more than offset the advantages and the com- 
mercially successful machine is yet to be found. 

The writer has tested out various types of rotary steam engines, and 
the best has been but a poor competitor of the ordinary grade of 
reciprocating mechanism. 

One of the most successful rotary engines is illustrated in Fig. 231. 
The device consists essentially of two rotors in rolling contact, the 
upper one containing a recess which serves as a steam inlet and allows 
the piston on the lower rotor to pass, while the lower one contains the 



STEAM ENGINES 



383 



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Fig. 233. Cost of Simple High-speed Engines. 



384 



STEAM POWER PLANT ENGINEERING 





























































































































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Fig. 234. Cost of High-speed Compound Engines. 



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Fig. 235. Cost of Low-speed Engines, Simple and Compound. 



STEAM ENGINES 385 

piston and transmits the power to the shaft. In fundamental prin- 
ciple it is not unlike many other rotary engines in that the power is 
applied directly to the shaft by the expansion of steam behind a rotary 
piston. The synchronous movement of the two rotors is maintained 
by means of two timing gears on the far side of the casing. The curves 
in Fig. 232 are based upon the tests made by Professor Pryor of Stevens 
Institute of a 20-horse-power engine of this design, initial pressure 
150 pounds gauge, atmospheric exhaust, steam dry and saturated. 

213. Cost of Engines. — In general the cost of engines per horse 
power diminishes as the size increases, but is of course governed by the 
style and workmanship. Average figures may be expressed as follows 
(Engr. U. S., Nov. 15, 1902, p. 750): 

Simple high-speed engines Cost in dollars = 300 +8 X horse power 

Setting, high-speed engines Cost in dollars = 60 + 0.75 X horse power 

Compound high-speed engines Cost in dollars = 1000 + 15 X horse power 

Simple low-speed engines Cost in dollars = 1000 + 10 X horse power 

Compound low-speed engines Cost in dollars = 2000 + 13 X horse power 

Setting, low-speed engines Cost in dollars = 500 +1.3 X horse power 

These equations were deduced from the curves in Figs. 233 to 235, 
which were plotted from the actual costs of a large number of engines. 

Rules for testing steam engines. — See Appendix C. 

For a complete description of modern types of piston engines see 
Prac. Engr. U. S., Jan. 1, 1913. 



CHAPTER X. 



STEAM TURBINES. 



214. Classification. — The development of the steam turbine during 
the past decade has been truly remarkable. So rapid has been the 
growth that many turbines representative of the best practice seven 
years ago are virtually obsolete to-day. Because of the almost radical 
changes from year to year it is practically impossible to keep the de- 
scriptive features of a textbook strictly in accord with current practice, 
and the subject matter must necessarily be of a general nature. 

Steam turbines are now being Used for driving alternating-current 
generators, turbo-compressors, pumps, blowers and marine propellers, 
and, by means of gearing to furnish power for reciprocating air com- 
pressors, rolling mills and other classes of slow-speed machinery. 
Although the reciprocating engine will probably continue to be an im- 
portant factor in the power world for years to come, its field of useful- 
ness is being gradually limited by the steam turbine. The steam turbine 
has found favor chiefly on account of its low first cost, low maintenance 
cost, small floor-space requirements and low cost of attendance. 

A general classification of steam turbines is unsatisfactory because 
of the overlapping of the various groups, and the following chart is 
offered merely as a guide in arranging a few well-known turbines ac- 
cording to the fundamental principles involved in their operation. 

f Single 

Velocity. 



Steam 
Turbines 



Impulse 



De Laval. 



Multi-velocity 
Stage. 



Single-velocity 
Stage. 



r Terry. 
J Sturtevant. 
] Curtis (SmaU Type). 
L Riedler-Stumpf . 

r Kerr. 

De Laval. 
-I Wilkinson. 

Rateau. 

Zoelly. 



Single- 
y pressure 
! Stage. 



Multi-velocity j Cmt ^ 
Stage. I 



Reaction. 

Combined 
Impulse and 
Reaction. 



Multi-velocity 
Stage. 

Multi-velocity 
Stage. 

386 



f Westinghouse- 

Parsons. 
I Allis-Chalmers- 
L Parsons. 

Westinghouse 
Double Flow. 



Multi- 

!► pressure 

Stage. 



STEAM TURBINES 387 

As shown in the preceding chart, all turbines may be divided into 
three general classes, (1) impulse, (2) reaction, and (3) combined impulse 
and reaction, though strictly speaking all turbines depend more or less 
upon both impulse and reaction for their operation. 

Impulse Type: 

In the impulse type the steam is expanded by suitable means and the 
heat given up by the pressure drop imparts velocity to the jet itself. 
The jet impinges against the vanes of a rotating wheel and gives up its 
kinetic energy to the wheel. If the entire pressure drop takes place in 
one set of nozzles and the resulting jet is directed against a single wheel 
the turbine is classified with the single-stage single-velocity group. The 
velocity of the jet is very high, from 2000 to 4000 feet per second, and 
for satisfactory economy the peripheral velocity of the wheel must also 
be very high, from 700 to 1400 feet per second. The De Laval single- 
stage turbine is the best-known example of this group. 

If the entire pressure drop takes place in a single set of nozzles and 
a single wheel is to be used at a comparatively low speed satisfactory 
economy may be effected by compounding the velocity. That is, the 
jet issuing from the nozzle at a very high velocity is reflected back and 
forth from the vanes on the rotor to a series of fixed reversing buckets 
until all of the available kinetic energy of the jet has been imparted 
to the wheel. The Terry single-stage turbine is representative of this 
group. 

Low peripheral velocity and high efficiency may be obtained by 
pressure compounding; that is, expansion takes place in a series of 
successive nozzles instead of one nozzle. Only a part of the available 
heat energy is converted into kinetic energy in each set of nozzles. For 
each set of fixed nozzles there is a corresponding rotor. This type of 
turbine is to all intents and purposes a series of single-velocity impulse 
turbines placed side by side. The Kerr turbine is representative of 
this group. 

By compounding both velocity and pressure we have the multi-velocity 
and pressure type of which the Curtis turbine is the best-known example. 

Reaction Type: 

In the reaction type the conversion of potential to kinetic energy 
takes place in the moving blades as well as in the fixed blades. Only a 
very small portion of the heat energy imparts velocity in the first set of 
fixed blades or nozzles. The jet issuing from this set of nozzles impinges 
against the first set of moving blades and imparts its kinetic energy to 
the rotor by impulse. The adjacent moving blades are proportioned so 
that partial expansion takes place within them and the resulting increase 



388 STEAM POWER PLANT ENGINEERING 

in velocity exerts a reaction which still further accelerates the rotor. The 
expansion is very gradual and a large number of alternately fixed and 
revolving blades are necessary to effect complete expansion. Because 
of the small pressure drop in each stage low peripheral velocities are 
possible with high over-all efficiency. The Westinghouse and Allis- 
Chalmers designs of the Parsons turbine are the best-known examples 
of this type. 

Combined Impulse and Reaction Type: 

In this class the high-pressure elements are of the impulse type and 
the low-pressure elements of the reaction type. The Westinghouse- 
Parsons double-flow high-pressure turbine is typical of this class and 
is virtually a combination of the Curtis and Parsons designs. Several 
European impulse turbines as recently designed are fitted with reaction 
blades adjacent to the nozzles, showing the tendency to merge the 
different fundamental types. 

Turbines may be classified according to the service for which they 
are intended, as 

High-pressure non-condensing, 

High-pressure condensing, 

Low-pressure, 

Mixed-pressure, 

Bleeder. 

Each of these types is discussed later on in the chapter. 

Recent Developments in Steam Turbine Practice: Mech. Engr., Jan. 26, 1912. 

The Present State of Development of Large Steam Turbines : Jour .A.S.M.E.,May,1912. 

The Steam Turbine: Engng., Dec. 29, 1911. 

Status of the Small Steam Turbine: Power, Jan. 2, 1912. 

215. General Elementary Theory. — A given weight of steam at a 
given pressure and temperature occupies a certain known volume and 
contains a known amount of heat energy. If the steam is permitted to 
expand to a lower pressure without receiving additional heat or giving 
up heat to surrounding bodies it is capable of doing a certain amount 
of work which will be the same whether the expansion takes place in 
the cylinder of a reciprocating piston engine, a rotary piston engine, or 
the nozzles and blades of a steam turbine. 
Let W = weight of steam, lbs. per sec. 

E = energy given up by 1 lb. of steam, ft. -lbs. 
Pi = initial pressure, lbs. per sq. in. abs. 
P n = final pressure, lbs. per sq. in. abs. 
Hi = initial heat content per lb., B.t.u. 
H n = final heat content per lb., B.t.u. 



STEAM TURBINES 389 

Then the heat drop, or heat available for doing useful work, is 

W (Hi - H n ) B.t.u. (124) 

If the steam expands against a resistance, as, for example, the piston 
of a reciprocating engine, the energy given up in forcing the piston 
forward may be expressed 

E 1 = 777.5 W (#! - H n ) ft.-lbs. (125) 

If the steam expands within a perfect nozzle the energy will be given 
up in imparting velocity to the steam itself, thus: 

#2 = W^ ft.-lbs. (126) 

in which 

Vi = velocity of the jet in feet per second. 

If the velocity of the jet is retarded to V n feet per second,' as by 

placing a series of vanes in its path, then the energy given up to the 

vanes (neglecting all losses) is 

Fi 2 — V 2 
E = W x _ n • (127) 

2g 

If the jet is brought to rest by the vanes (neglecting all losses), then 
V n = and the energy given up is 

*-wJ£- (128) 

777.5 W (ffi - H„) = W^, 



But Ei = E 3 . Hence, 
from which 



Vi = 223.8 VH X - H n * (129) 

If there are n pressure stages, then the theoretical stage velocity is 

Vi f = 223.8 \/ Hl ~ Hn • (130) 

The jet issuing from the nozzle is capable of exerting an impulse 
equal to F upon any object in its path, thus: 

F = Illi lbs. (131) 

9 

If A = the area of cross section of the jet in square feet, and y = 
weight of steam, pounds per cubic foot, then W = yAV h or 

F = Y^L 2 lbs> (132) 

9 

* For most purposes it is sufficiently accurate to make 223.8 = 224. 



390 STEAM POWER PLANT ENGINEERING 

The reaction, R, of the jet against the nozzles is equal in value and 
opposite in direction to the impulse, or 

R = F = — i = ^^-i- . (133) 

9 9 

The theoretical horse power developed by a jet of steam flowing at 
the rate of one pound per second may be expressed 

E y.2 _ y 2 

H - p -=55cT2TxW ■ (134) 

in which 

Vi = initial velocity of the jet, ft. per sec. 
V n = final velocity of the jet, ft. per sec. 

Steam consumption per horse-power hour: 

Heat consumption, B.t.u. per horse power, per minute: 

60 ' (136) 

in which 

q n = heat of the liquid at pressure P n . 

Impulse efficiency of the jet = equation (127) -s- equation (128). 

7,2-72 
Ei= y/ n ' (137) 

Thermal efficiency (Rankine Cycle) : 

E= Hi-H n 
Hi — q n 
Efficiency ratio or "kinetic" efficiency: 

E = W.iH, - H n ) ■ (139) 

Equations (124) to (139) are general and are applicable to all turbines 
of whatever make. 

The more important types of turbines will be discussed separately 
and an application of above equations will be made in each specific case. 

Heat Drop in Steam Turbines: Trans. A.S.M.E., Vol. 33, p. 325, 1911; Engr., 
Mar. 8, 1912. 

216. The De Laval Turbine. — Fig. 236 shows a section through a 
De Laval steam turbine and gear case and illustrates the principles of 
the single-stage "impulse" type. The turbine proper, to the right of 
the figure, consists of a high-carbon steel disk C fitted at the periphery 
with a single row of drop-forged steel blades and inclosed in a cast-steel 



STEAM TURBINES 



391 




392 



STEAM POWER PLANT ENGINEERING 



casing. The disk is secured to a light flexible shaft and is of such a 
cross section that the radial and tangential stresses throughout its mass 
are of constant value. A flexible shaft is employed which allows the 
wheel to assume its proper center of rotation and thus to operate like 
a truly balanced rotating body.* The shaft is supported by three 

bearings, P, K, and /. / is self- 
aligning and carries the greater 
part of the weight of the disk. 
K is a flexible bearing, entirely 
free to oscillate with the shaft, 
and its only function is to seal 
the wheel casing against leakage. 
The power is transmitted 
through a steel helical pinion K' 
mounted on the extension of the 
turbine shaft E, to two large 
gears M, M at a reduction in 
speed of about 10 to 1. The 
blades, Fig. 237, are made with a bulb shank and fitted in slots milled in 
the rim of the wheel. The flanges, at the outer end of the blades, are 
brought in contact with each other and calked so as to form a continu- 
ous ring. The inlet and outlet angles of the blades are made alike and 
are 32 degrees for smaller sizes and 36 degrees for larger sizes. 



v//////////////////////////////////^ >^ 







Fig. 237. De Laval Blades. 




Fig. 238. De Laval Nozzle. 

The operation is as follows: Steam enters the steam chest D, Figs. 
236 and 237, through the governor (shown in detail in Fig. 239) and 
is distributed to the various adjustable nozzles, varying in number 
from 1 to 15 according to the size of turbine. In the earlier types 
the nozzles were uniformly distributed around the circumference, but 
in the later types are arranged in groups. As illustrated in Fig. 238 

* The shaft diameter for a 100-horse-power turbine is but 1 inch and for a 300- 
horse-power turbine approximately l^f inches. 



STEAM TURBINES 



393 



the nozzles are placed at an angle of 20 degrees with the plane of the 
disk. The steam is expanded adiabatically in the nozzles to the existing 
back pressure before it impinges at high velocity against the blades. 
After giving up its energy the steam passes into chamber W, Fig. 236, 
and out through the exhaust opening. Fig. 239 gives the details of 
the governor and vacuum valves. Two weights B are pivoted on knife 
edges A with hardened pins C bearing on the spring D. E is the governor 
body, fitted in the end of the gear-wheel shaft K, and has seats milled 
for the knife edges A. The spring seat D is held against pins A by 




Fig. 239. De Laval Governor. 

spiral concentric springs, the tension on which is adjusted by a milled 
nut /. When the speed exceeds the normal, centrifugal force causes 
the weights to fly outward and overcome the resistance of the springs. 
This pushes pin G against bell crank L, which in turn closes the double- 
seated valve, thus throttling the supply of steam. To prevent racing 
in case the load is suddenly removed the vacuum valve T is added to 
the governor mechanism. Its operation is as follows: The governor 
pin G actuates the plunger H under normal conditions without moving 
the plunger relative to the bell crank. In case the load is suddenly 
removed, centrifugal force pushes pin G against bell crank L until it 
reaches its extreme position and- the valve is nearly closed and little 
steam enters the turbine. If this does not check the speed, plunger G 



394 STEAM POWER PLANT ENGINEERING 

overcomes the resistance of spring M, and H moves relative to L, and 
its adjustable projection presses against valve stem T and allows air 
to rush into the turbine chamber through passage P. 

The power of the turbine depends upon the number of nozzles in 
action, and these can be opened or closed by a hand wheel on each. 
Each nozzle performs its function as perfectly when operating alone as 
when operating in conjunction with others. 

De Laval turbines of the single-stage geared type are made in sizes 
ranging from 17 to 700 horse power, condensing and non-condensing, 
and are designed to regulate within an extreme variation of 2 per cent 
from no load to full load. 

The speeds vary from 10,600 r.p.m. for the largest size to 30,000 r.p.m. 
for the smallest, the gearing reducing these to 900 and 3000 r.p.m., 
respectively, at the shaft. 

The diameter of the wheel varies from 4 inches in the smallest tur- 
bine to 30 inches in the largest, thus giving peripheral velocities of from 
520 to 1310 feet per second. 

In addition to the single-stage geared type the De Laval company 
manufacture the single-stage gearless, two-stage gearless, and multi- 
stage gearless turbine. The latter are constructed in sizes up to 5000 
kilowatts and are described in paragraph 221. 

217. Elementary Theory. — De Laval Single-stage Turbine. — The 
maximum theoretical power developed by a jet of steam flowing through 
a nozzle is dependent only upon the weight of steam flowing per unit of 
time and the initial velocity. Therefore the higher the initial velocity 
for a given rate of flow the greater will be the power developed and the 
higher the efficiency. 

The maximum weight of steam discharged through a nozzle of any 
shape and for a given initial pressure is determined by the area of the 
narrowest cross section or throat. 

To obtain the maximum velocity at the exit or mouth, for a given rate 
of flow, the nozzle should be proportioned so that expansion to the 
external pressure into which the nozzle delivers shall take place within 
the nozzle itself. If expansion in the nozzle is incomplete, sound waves 
will be produced and there will be irregular action and loss of energy. 
On the other hand, if expansion in the nozzle is carried below that of 
the external pressure at the mouth, sound waves will be produced with 
subsequent loss of energy even greater than in the former case. 

Experimental and mathematical investigations indicate that the pres- 
sure at the narrowest section of an orifice or the throat of a nozzle 
through which steam is flowing falls to approximately 0.58 of the initial 
absolute pressure (with resultant velocity of about 1400 to 1500 feet per 



STEAM TURBINES 



395 



second) and any farther fall in pressure must take place beyond the 
narrowest section. Thus for back pressures greater than 0.58 of the 
initial (conveniently taken as f ), maximum exit velocity may be ob- 
tained from orifices or nozzles of uniform cross section or with sides 
convergent. For back pressure less than 0.58 of the initial the nozzle 
must first converge from inlet to throat and then diverge from throat to 
mouth in order to obtain maximum velocity. Without the divergent 
portion of the nozzle the jet will begin to spread after passing the throat, 
and its energy will be given up in directions other than that of the 
original jet. 

Fig. 240 shows a section through a theoretically proportioned ex- 
panding nozzle. The cross section of the tube at any point n may be 




^Mw/mMmm 



Fig. 240. Theoretically Proportioned Expanding Nozzle. 

calculated by means of equation 



A n = 



WS n 

' n 



(140) 



in which 

A n = area in square feet. 

W = maximum weight of steam discharged, pounds per second. 

S n = specific volume of the steam at pressure P n . 

For saturated steam S n = x n u n , 
in which x n = quality of steam at pressure P n after adiabatic expansion 
from pressure Pi. 
u n = specific volume of saturated steam at pressure P n . 
For superheated steam, see Mollier diagram, Appendix L. 
V n = velocity of the jet, feet per second. 
V n may be determined from equation (129): 
V n = 223.8 VH! - H n . 

By substituting H n = heat content corresponding to pressure 
P n = 0.58 Pi in equations (129) and (140) the area at the throat may be 
readily determined. The cross-sectional area for other points in the 



396 



STEAM POWER PLANT ENGINEERING 



tube may be determined in a similar manner by. assigning values of 
H n corresponding to the various pressures. 

In case of a perfect nozzle Hi — H n represents the heat given up 
toward producing velocity by adiabatic expansion from pressure Pi to 
P n . In the actual nozzle the frictional resistance of the tube serves to 
increase its dryness fraction, but in doing so it decreases the amount of 
energy the steam is capable of giving up towards increasing its own 



THEORETICAL DESIGN OF A DIVERGENT NOZZLE 



5000 300 



4500 



4000 



190? 

180 

170 

160 

150 



3500 U 140 

•o 5, 

a & 

M 130 

9, *-* 

f 3000 a 120 

I i 

<u o 

g 2500^ 100 

a B 
i i 9 

1 2000 | 80 

^ a 

£ 1500 g 60 



QQ 



50 



©10003 40 



500 



1 
























































\ 




























\ 




























\ 
































































































,4 




























**/ 


























A / 




























Q 


£afi 






























t£l 














































































































































^ e ' 


&s 
























3 


*s^« 


























^ 


-A 


>Vn 





















































B 


6 


5 

a 








© 




a 




H 






5 


o 


100 


72 




O 


90 


4^ 




K 




CjH 


4 


0) 

If) 


80 






a 

S3 








70 


"3 


3 


Q 


60, 


a 



40,000 80,000 120,000 160,000 200,000 240,000 280 v 0OO 
Kinetic Energy of the Jet, in Foot Pounds 

Fig. 241. 

velocity. If y one-hundredths of the heat H 1 - H n is utilized in over- 
coming frictional resistance, then the resulting velocity will be 

V = 223.8 V(l - y) (H, - H n ). (141) 

The quality of the steam after expanding to P n against the resistance 
will be higher by an amount 

I n = increase in quality = — > (142) 

Tu- 
rn which 



r n = heat of vaporization at pressure P n . 



STEAM TURBINES 



397 



The curves in Fig. 241, calculated by means of equations (129) and 
(140), show the relationship between velocity, quality, pressure, and 
kinetic energy for all points in a theoretically perfect nozzle expanding 
one pound of dry steam per second from an initial absolute pressure of 
190 pounds to a condenser pressure of one pound. 

The curves in Fig. 242 are based upon the experiments of Gutermuth 
(Zeit. d. Ver. Ingr., Jan, 16, 1904) and show the effect of a few shapes 
of nozzles and orifices on the actual weight of steam discharged for 
various rates of initial and final pressures, the smallest section of the 
tube remaining constant. 



.06 



.05 



.04 



| .02 



.01 



1 


































■ 




































4 






























V 






















s 


\ 














^1 
























>3 


\ 
















*N 






































\ 


1 






































\ 








, 


















\\ 




P— 

1 









P 




\ 




A 


'£%& 
















, 








\ 






_^ 




3 










\ 




p r 


..... 


P 


P- 

1 


r~ 


P, 






\ 






































\ 






1 
















- 






1 






2 












P— 


-v 1 


P 










r 




P=132 Lb. Per Sq. In. Absolute 












4 












A 


rea 


3fOl 


ifi.ce 


0.0 


J55S 


q.In 


. 

























06 



05 



.04, 



.01 



.1 .2 .3 .4 



.7 .8 .9 1.0 .1 .2 



.4 .5 .6 .7 .8 .9 1.0 



Eatio— P-J-P 



Fig. 242. Flow of Steam through Nozzles. 

The nozzles of most commercial types of steam turbines are made 
with straight sides as in Fig. 238, so that only the area at the mouth 
need be determined in addition to that at the throat in order to lay out 
the shape of the tube. 

Equations (129) and (140) are general and are applicable to steam 
of any quality, wet, dry, or superheated. For steam initially dry and 
saturated Napier's rule offers a simple means of determining the area 
at the throat, thus: 

W = ^forP n =or<|p 1 , (143) 



W = 0.029 A VPn(Pi - P n ) for P n > | P h (144) 

in wnicn 

W = maximum weight of steam discharged, pounds per second; 

A = area at the throat, square inches; 

Pi = absolute initial pressure, pounds per square inch; 

P n = absolute back pressure, pounds per square inch. 



398 



STEAM POWER PLANT ENGINEERING 



Moyer ("The Steam Turbine/' 1st Edition, p. 40) states that the ratio 
of the area of a correctly proportioned nozzle at the throat A to the 
area at any point A n is very nearly proportional to the ratio of the 
pressure at point A n to the initial pressure, or 

Ao = Pi. m 



(145) 



The entrance to the tube is rounded by any convenient curve. 
The length of the tube may be roughly approximated by the follow- 
ing formula: 

L = Vl5A , (146) 

in which 

L = length between the throat and mouth, in inches; 
Aq = area at the throat, square inches. 

Practice shows that the cross section of a nozzle, whether circular, 
elliptical, square, or rectangular (the latter with rounded corners), has 
very little influence on the efficiency provided the inner surfaces are 
smooth and the ratio of the area at the throat to that of the mouth is 
correctly proportioned. The velocity efficiency of a properly propor- 
tioned nozzle with straight sides is about 95 to 97 per cent, correspond- 
ing to an energy efficiency of 92 to 94 per cent, so that it is not con- 
sidered worth while to attempt to follow the more difficult exact curves. 

Example: — Find the smallest cross section of a frictionless conically 
divergent nozzle for expanding one pound of steam per second from an 
.absolute initial pressure of 190 pounds to an absolute back pressure of 
2 pounds and find six intermediate cross sections where the pressures 
will be 70, 30, 14.7, 8, 4, and 2 lbs. respectively. Compare the velocity 
and energy of the jet issuing from this nozzle with those of an actual 
nozzle in which 10 per cent of the heat energy is lost in friction. 

From steam and entropy tables we find the values of H, x, u, for 
absolute pressures corresponding to 190, 0.58 X 190 = 110, 70, 30, etc., 
lbs. per square inch as follows (theoretical nozzle) : 





H. 


X. 


u. 


S = xu. 


P l = 190 


1197.3 


1.00 


2.405 


2.394 


P 2 = no* 


1152.6 


0.960 


4.047 


3.878 


P., = 70 


1117.9 


0.932 


6.199 


5.775 


P 4 = 30 


1057.2 


0.887 


13.75 


12.27 


P 5 = 14-7 


1011.3 


0.857 


26.78 


22.95 


P«= 8 


947.8 


0.834 


47.26 


39.29 


P 7 = 4 


935.6 


0.810 


90.4 


73.2 


P 8 = 2 


899.3 


0.788 


173.1 


137.0 



* P 2 = 0.58 Pi ( = pressure at throat). 



STEAM TURBINES 



399 



If entropy tables or charts are not available, values Hi to H 8 and 
Xi to x 8 must be calculated. 

The different quantities for the theoretical nozzle will be calculated 
for the exit pressure P n = P 8 = 2 lbs. per sq. in. absolute. 

V 8 = 223.8 Vff x - ff 8 

= 223.8 V1197.3 -899.3 
= 3865 feet per second. 
E 8 = 778 (Hi - H 8 ) 

= 778 (1197.3 - 899.3) 
= 232,000 foot-pounds. 
WS 
V 
1 X 137 



A 8 =^ 



3865 
= 0.0353 square foot. 

= V /(^ij A=1 3.56VJ 



F 8 = 



13.56 Vo.0353 
2.54 inches. 
WV* 



9 
3865 



32.2 

120 pounds. 



THEORETICAL NOZZLE. 



Quantity < 


V 
Ft. per Sec. 


E 
Ft .-Lbs. 


A 

Sq. Ft. 


d 
Inches. 


F 
Pounds. 




(73) 


(72) 


(76c) 




(74) 




Pressures 


110 
70 
30 
14.7 

8 

4 

2 


1,496 
1,995 
2,650 
3,053 
3,339 
3,624 
3,865 


34,767 
61,853 
107,485 
144,742 
173,207 
203,968 
232,000 


.00259 

.00269 

.00461 

.00745 

.0119 

.0202 

.0353 


0.693 

0.702 

0.919 

1.1 

1.46 

1.92 

2.54 


46.4 

61.98 

82.3 

94.8 

103.7 

112.5 

120 . 



In the actual nozzle these values will be modified because of the 
frictional losses. Thus, for P n = 2 lbs., 

T^8 = 223.8 V(l - y )(H 1 -H 8 ) 



= 223.8 V(l -0.1) (1197.3 
= 3667 ft. per sec. 



899.3) 



400 



STEAM POWER PLANT ENGINEERING 



E 8 = 778(1 -0.1)(1197.3 



X& = X 8 ~\r 1 8 = X S 



y(H } 



899.3) 



208,800 ft.-lbs. 



Ts 



= 0.788 + 



0.1 (1197.3 - 899.3) 



1021 



= 0.788 + 0.029 
= 0.817. 

Wxs'us 



from which 



_ 0.817 X 173.1 

3667 
= 0.0386 sq. ft., 

d s = 2.66 in. 



These various factors for all given pressures have been calculated in 
a similar manner and are as follows: 



ACTUAL NOZZLE. 



Quantities -j 


V 
Ft. per Sec. 


E 
Ft .-Lbs. 


x'. 


A 
Sq. Ft. 


d 
Inches. 


F 

Ft .-Lbs. 




110 


1,420 


31,317 


.9658 


.00275 


0.711 


44.1 




70 


1,893 


55,632 


.9414 


.00286 


0.723 


58.8 




30 


2,515 


98,257 


.9026 


.00493 


0.951 


78.12 


Pressures ■ 


14.7 


2,894 


130,050 


.876 


.0080 


1.2 


98.8 




8 


3,168 


155,858 


.856 


.0127 


1.53 


98.4 




4 


3,438 


183,581 


.836 • 


.0220 


2.01 


106.8 




2 


3,667 


208,800 


.817 


.0386 


2.66 


114.0 



Many of these values may be determined directly from the Mollier 
or total heat-entropy diagram as described in Appendix L; in fact, 
the Mollier diagram has to all intents and purposes supplanted the 
steam tables in this connection. For superheated steam the diagram 
is extremely useful in avoiding laborious calculations. 

Fig. 243 gives a diagrammatic arrangement of the blades in a single- 
stage De Laval turbine. The nozzle directs the steam against the blades 
with absolute velocity V\ and at an angle a with the plane of the wheel 
XX. Since the wheel is moving at a velocity of u feet per second, the 
velocity vi of the steam relative to the wheel is the resultant of V\ and 
u. The angle & between v\ and XX will be the proper blade angle at 
entrance. If the blade curve makes this angle with the direction of 
motion of the wheel no shock will be experienced when the steam enters 



STEAM TURBINES 



401 



the blades. For convenience in construction the exit angle ft is made 
the same as the entrance angle ft. Neglecting frictional losses in the 
blade channels the relative exit velocity will be v 2 = vi, and the absolute 
velocity V 2 is the resultant of v 2 and u. The impulse exerted by the 

W 

jet in striking the vanes is — v h and its component in the direction of 

W W 

motion is — Vi cos ft = — (Vi cos a — u). As the jet leaves the vanes 



W 

the impulse is v 2 cos ft 



W 



(T 2 cos 7 + u). 



W=1250 




Fig. 243. Velocity Diagram. Ideal Single-stage Impulse Turbine. 

The total pressure acting on the vanes, or the actual driving impulse, is 
P = — \Vi cos a — u — \ — (V 2 cos y + u) > 



W 

= — (Vi cos a + V 2 cos 7). 



(147) 



Equation (147) may also be expressed 
.. W 



2 (FiCOSo: — u). 



(148) 



The resultant axial force or end thrust is 

W 

F = — (7i sin a - V 2 sin 7). (149) 

Evidently if a = 7 and Vi = V 2 there will be no end thrust, since 
Vi sin a — V 2 sin 7 will be zero. 
The work done is 



W 

Pu = — u (V\ cos a-\- V 2 cos 7), 



(150) 



402 STEAM POWER PLANT ENGINEERING 

or, using equation (148) in place of (147), 

W 

Pu = — •2u (Vi cos a — u) 

a 

w 

= — .2(wF lC osa-w 2 ). (151) 

y 

By making the first derivative equal to zero 

— < — 2 (uVi cos a — u 2 ) > = Vi cos a — 2 u = 0, 
du ( g ) 

or w = |7i cos a. 

That is, for any nozzle angle a the work done, Pu, has its greatest 
value when u = J Vi cos a or y = 90 degrees, whence 

p w = IF ^ cos 2 a. (152) 

2# 

The work for any initial velocity Y\ becomes a maximum when a = 
and w = \ Vi. This condition can only occur for a complete reversal of 
jet and zero final velocity. Substitute a = and u — J V\ in equation 

(151). 

WVi 2 
Pu = —x — , which is necessarily the same as equation (128). 

In the actual turbine the various velocities will be less than those as 
obtained on account of the frictional resistance in the blades, and the 
velocity diagram should be modified accordingly. 

Example: Lay out the blades (theoretical and actual) for the nozzle 
in the preceding example, assuming that the jet impinges against the 
wheel at an angle of 20 degrees and that the peripheral velocity is 
1250 feet per second. 

Theoretical Case: 

Lay off Vi = 3865 feet per second in direction and amount as shown 
in Fig. 243 and combine it with u = 1250 feet per second; this gives 
Vi, the relative entrance velocity, as 2725 feet per second, and /3, the 
entrance angle, as 29 degrees. 

Lay off v 2 = Vi at an angle /3 2 = ft and combine with u; this gives 
V 2 , the absolute exit velocity, as 1740 feet per second. 

The theoretical energy available for doing work is 

W 

= ^j— (3865 2 - 1740 2 ) = 185,000 foot-pounds. 



STEAM TURBINES 



403 



The difference between 232,000 and 185,000 = 47,000 foot-pounds is 
evidently the kinetic energy lost in the exhaust due to the exit velocity. 
The pressure exerted by the steam on the buckets is 

W 

P = — (Vi cos a + F 2 cos 7) 
9 

=i- 2 (3865 X 0.9397 + 1740 X 0.65166) 
t= 148 pounds. 
The theoretical impulse efficiency is 

ti 2 -F 2 



3865' 



1740 2 



vs 



3865' 



= 0.797. 




U = 1250 

Fig. 244. Velocity Diagram as Modified by Friction Losses. 

The theoretical horse power per pound of steam flowing per second is 

185,000 



H.P. 



550 



= 336. 



Theoretical steam consumption per horse-power hour is 
3600 



336 



= 10.7 pounds. 



Actual Case: 

Proceed as in the theoretical case, using the actual absolute velocity 
Vi = 3865 Vl - y = 3865 Vl - 0.10 = 3667 feet per second in place 
of the theoretical value Vi = 3865. Lay off Vi = 3667 at an angle of 
20 degrees as before and combine with u = 1250, Fig. 244. 

The resultant vi = 2530 is the velocity of the jet relative to the wheel, 
and'the entrance angle /3 is found to be 29.7 degrees. The relative exit 
velocity v% will be less than vi because of the blade friction. 



404 STEAM POWER PLANT ENGINEERING 

Assume the loss of energy from this cause to be 14 per cent; then, 
since the velocity varies as the square root of the energy, 

v 2 = vi Vl -0 (153) 

= 2530 Vl - 0.14 
= 2346 feet per second. 

The resulting absolute velocity V 2 is found from the diagram to be 
T 2 = 1405 feet per second. 
Since the loss of energy in the nozzle is 

V '-%; V)V ' , (154) 

and that in the blade 

«-%-»« (155) 

the remaining energy, deducting both losses in the nozzle and the 
blades, is 

^ (TV - yVf - <j> vS - IV) (156) 

= -^- (3865 2 - 0.1 X 3865 2 - 0.14 X 2530 2 - 1405 2 ) 
= 164,200. 

The losses due to windage, leakage past the buckets and mechanical 
friction must be deducted from these figures to give the actual energy 
available for doing useful work. Assuming a loss of 15 per cent due to 
this cause, the work delivered is 

0.85 X 164,200 = 139,570 foot-pounds. 

The efficiency in the ideal case was found to be 0.797 and the avail- 
able energy 185,000 foot-pounds. 

The efficiency, deducting the loss due to friction, etc., is 

!-tS xo - 797 = - 60 - 

The horse power delivered is 

139,570 _ 
~~550~ " ^ 4 ' 

Steam consumption per horse-power hour is 

3600 1AO A 

-2gj = 14.2 pounds. 

The heat consumption, B.t.u. per horse power, per minute is 
14.2 (1197.3 - 94) 

60 JbU ' 



STEAM TURBINES 



405 



Assuming the revolutions per minute to be 10,000, the mean diameter 
of the wheel to give a peripheral velocity of 1250 feet per second is 
1250 X 60 . on - , OCft . , 
10,000 X 3.14 = 2 ' 39 feet ' ° r 28 ' 6 mcheS ' 

The determination of the height and width of vanes, clearance be- 
tween nozzles and blades, etc., are beyond the scope of this work and 
the reader is referred to the accompanying bibliography. 

Blade Design for De Laval Turbines: Moyer, "Steam Turbine," Chap. IV; Power, 
Mar. 17, 1908, p. 391. 

Flow of Steam through Nozzles: Jour. A.S.M.E., Mid. Nov., 1909, April, 1910, 
p. 537; Engineering, Feb. 2, 1906; Engr., Lond., Dec. 22, 1905; Eng. Rec., Oct. 26, 
1901; Power, May, 1905; Eng. News, Sept. 19, 1905, p. 204. 

Design of Turbine Disks: Engr., Lond., Jan. 8, 1904, p. 34, May 13, 1904, p. 481. 

Turbine Losses and their Study: Jour. El. Power and Gas, March 9, 1912. 

Critical Velocity of Shafting: Jour. A.S.M.E., June, 1910, p. 1060; Power, Sept., 
1903, p. 484. 




Fig. 245. Section through Single-stage Terry Steam Turbine. 

218. Terry Turbine. — Fig. 245 shows a section through a single- 
stage Terry turbine, illustrating an application of the single-stage im- 
pulse type with two or more velocity stages. This " compounding" of 
the velocity permits of much lower peripheral velocities than with the 
single-velocity type. The rotor, a single wheel consisting of two steel 
disks held together by bolts over a steel center, is fitted at its periphery 



406 



STEAM POWER PLANT ENGINEERING 



with pressed-steel buckets of semi-circular cross section. The inner 
surface of the casing is fitted with a series of gun-metal reversing buckets 
arranged in groups, each group being supplied with a separate nozzle. 
The steam issuing from nozzle N, at very high velocity, Fig. 246, strikes 
one of the buckets, B, on the wheel, and since the velocity of the buck- 
ets is comparatively low, is reversed in direction and directed into the 
first one of the reversing chambers. The chamber redirects the jet 
against the wheel, from which it is again deflected; this is repeated 
four or more times until the available energy has been absorbed by 
the rotor. Terry turbines are made in a number of sizes varying from 
5 to 800 horse power, and operate at speeds varying from 210 feet per 
second in the smaller machine to 260 feet per second in the larger. 




Fig. 246. Arrangement of Buckets and Reversing Chambers in a Terry Steam Turbine. 



These low speed limits compared with the speed of single-stage De Laval 
turbines are made possible by the application of the velocity-stage 
principle in the use of the reversing buckets. The rotor of the smaller 
machine is 12 inches in diameter and runs at 3800 r.p.m., and that of 
the larger, 48 inches, running at 1250 r.p.m. Since the flow of steam 
into and from the buckets is in the plane of the wheel there is no end 
thrust. 

Non-condensing Terry turbines are all of the single-stage type. The 
condensing units are of the two-stage type, the first stage expanding 
the steam from initial to atmospheric pressure and the second stage 
expanding the steam from atmospheric to condenser pressure. 

For a description of the Bliss, Dake, Sturtevant and Wilkinson 
steam turbines with results of tests see "Small Steam Turbines/ ' by 
G. A. Orrok, Jour. A.S.M.E., May, 1909, and contributed discussion, 
Sept., 1909. See also "The Development of the Small Steam Tur- 
bine," Eng. Mag., Dec, 1908, and Jan., 1909. 




STEAM TURBINES 



407 



219. Elementary Theory. — Terry Turbine. — Fig. 247 gives the 
theoretical velocity diagram for a single pressure stage Terry Turbine. 
Since the entire heat drop takes place in the nozzle the initial velocity 
of the jet OA is the same as with the single-stage De Laval turbine and 
may be calculated by means of equation (129) . OA represents the ab- 
solute velocity of the jet, OC the peripheral velocity and AOC the angle 
of the nozzle. CB is the component, parallel to the line of the jet, of 
the resultant of AO and OC. DC, in line with and equal in length to 
CB, combined with the peripheral velocity DE gives EC, the absolute 
velocity of the steam as it leaves the first set of rotating buckets. OiF, 
parallel to OA and equal in length to EC, represents the velocity of the 
steam as it enters the first stationary or reversing bucket. JG is the 




Fig. 247. Theoretical Velocity Diagram, Terry Turbine. 

component of the resultant of OiF and OiJ in line with the jet. The 
resultant IJ of HJ ( = JG) and HI represents the velocity of the steam 
as it leaves the first stationary bucket. This construction is repeated 
through all velocity stages. The final exit velocity of the steam as it 
issues from the moving buckets is WY. The energy converted into 
useful work is 

K= (OA 2 -WY 2 ). 

In the actual turbine friction losses would reduce the length of the 
velocity lines and increase the amount of energy rejected in the exhaust. 
The construction of the velocity diagram as modified by friction is 
similar to that described in paragraph 217, Fig. 240. 

220. Kerr Turbine. — Fig. 248 shows a longitudinal section through 
an eight-stage Kerr steam turbine illustrating the compound-pressure 
or multi-cellular group of the impulse type. The rotor consists of a 



408 



STEAM POWER PLANT ENGINEERING 




H 
S 

03 



H 



STEAM TURBINES 



409 




Fig. 249. Bucket Fastening, Kerr Turbine. 



series of steel disks, mounted on a rigid steel shaft. A series of drop- 
forged steel buckets is secured to the periphery and riveted in dove- 
tailed slots as shown in Fig. __ 

f ^7 — —r* —pj ^ J^F Tjf\ 

249. The tips of the buckets ] (Of; (Qi '<p'; .p!; '£)',' (Q'/'( 

are riveted to a shroud ring, — ™ — ~ — — "* ^' 

thereby insuring a rigid and 

positive spaced construction. 

The stator is made up of a 

number of arched cast-iron 

diaphragms with circular rims 

tongued and grooved, and 

bolted to steam-end and 

exhaust-end castings. The 

nozzles are formed by walls 

within the diaphragm and 

thin Monel metal vanes die- 
pressed into shape and cast 

into the diaphragm. One set 

of nozzles and one wheel constitute a stage and the expansion is usually 

carried out in from six to ten 
stages, depending upon the con- 
dition of operation. 

The operation is as follows: 
Steam enters the turbine through 
a double-beat balanced poppet 
valve, the stem of which is con- 
nected through levers to the 
governor, to the circular cored 
space H, H extending around the 
steam 1 1 end casting. ' ' This space 
acts as an equalizer and insures 
uniform admission to the first set 
of nozzles. Partial expansion 
takes place through the first set of 
nozzles and the kinetic energy is 
imparted to the rotor through the 
medium of the vanes. Steam 
leaves the buckets at a very low 
velocity and is again expanded 
through the second set of nozzles 

in the diaphragm. This process is repeated in each stage and exhaust 

steam leaves the turbine at 0. 




Fig. 250. 



Arrangement of Vanes and Nozzles, 
Kerr Turbine. 



410 



STEAM POWER PLANT ENGINEERING 



Fig. 251 illustrates the principles of the oil relay governor as applied 
to the larger sizes of turbines driving alternators. Referring to Fig. 251 : 
rotation of the turbine shaft is transmitted through worm gear and 
governor spindle to weights, W, W. Centrifugal force throws these 
weights outward about suspension points A and A f , overcoming the 
resistance of the spring. The movement of the spring is transmitted 
through lever L to relay plunger P and admits oil pressure (about 
30 pounds per square inch) to piston S and in this manner throttles 
admission valve V. Similarly, a downward movement of the relay 
plunger stem releases oil pressure and opens the admission valve. 




Fig. 251. Oil Relay Governor, Kerr Turbine. 

Floating lever L is connected to the admission valve stem through 
secondary lever M so that the movement of the steam valve returns 
the relay plunger to its central position. This equalizes the pressure on 
top and bottom of the main piston S and arrests its movement, thereby 
maintaining a fixed opening for a given speed. A suitable emergency 
valve automatically cuts off the steam supply in case the speed exceeds 
a predetermined amount. 

A spring-loaded governor of the centrifugal type mounted directly 
on the turbine shaft is used to control the smaller sizes of turbines. 

Kerr turbines are constructed horizontally and vertically and in 
various sizes ranging from 5 to 750 horse power, and are designed to 
operate all classes of pumps, blowers and generators. The rotative speed 



STEAM TURBINES 411 

varies from 2000 to 4000 r.p.m., depending upon the service for which 
the turbines are intended. 

221. The De Laval Multi-stage Turbine differs from the Kerr tur- 
bine only in mechanical details. It is of the multi-cellular type and 
is provided with a rigid shaft. The increase in the cross-sectional area 
of the passages required by the expansion of the steam as it proceeds 
through the turbine is effected by lengthening the blades, reducing the 
diameters of the wheels correspondingly and increasing the bore of the 
casing. (In the Kerr turbine the blades are lengthened and increased 
in width from the high-pressure to the low-pressure stages and the 
steam passages are increased in size but the outside diameter of the 
rotor remains the same.) The bearings are of rigid construction ar- 
ranged for water cooling. Labyrinth packing is used between stages 
and combined labyrinth and carbon-ring packing at the steam and 
exhaust ends of the casing. Air leakage into the turbine is prevented 
by introducing live steam between the two outer carbon rings. The 
governor is of the throttling type and is mounted upon a vertical shaft 
driven through worm gearing by the main turbine shaft. 

222. Elementary Theory. — Kerr Turbine. — In the Kerr turbine the 
diameters of the moving elements are uniform and hence the periph- 
eral velocities are the same. In the frictionless turbine the velocity 
issuing from each jet or the stage velocity should be the same. If 

there are n stages the heat drop per stage will- be - of the total heat 

drop. Since there are no friction losses in the ideal turbine the total 
heat drop is 

Hi — H n 
and the heat drop per stage 

Hi — H n 



The stage velocity or initial velocity of jet from each nozzle is 

V = 224,J Hl ~ Hn - 
V n 

The pressure, specific volume and quality of the steam in each stage 

TJ TT 

may be determined by subtracting — from the heat content of 

the preceding stage and finding the corresponding quantities from tem- 
perature-entropy tables or diagrams. 

Thus, an eight-stage turbine operating non-condensing at 190 pounds 
initial absolute pressure would show about the following conditions. 



412 



STEAM POWER PLANT ENGINEERING 



(All friction and leakage losses neglected and final velocity in each 
stage assumed to be zero.) 

Hi = 1197.3 B.t.u. per pound. 
H n = 1012.5 B.t.u. per pound. 
Total heat drop = H x - H n = 1197.3 - 1012.5 = 184.8. 
184.8 



Heat drop per stage 



8 



23.1 



Stage velocity = 224 V23.1 = 1080 feet per second. 



Stage. 


Heat Content. 


Pressure, Lbs. Abs. 


Quality, Per Cent. 


Specific Volume 
Cu. Ft. per Lb. 


Admission. 


1197.3 


190 


100 


2.41 


1 


1174.2 


145 


97.9 


3.04 


2 


1151.1 


109 


95.9 


3.93 


3 


1127.0 


80 


94.0 


5.14 


4 


1104.9 


58 


92.2 


6.77 


5 


1081.8 


42 


89.6 


8.96 


6 


1058.7 


30 


88.8 


12.07 


7 


1035.6 


21 


87.3 


16.33 


8 


1012.5 


Atmospheric 


85.8 


22.55 



In the actual turbine only 50 to 75 per cent of the heat theoretically 
available is transformed into useful work. A small portion is lost by 
gland leakage, radiation and bearing friction and the balance has been 
retransformed from kinetic energy into potential energy by eddying, 
fluid friction and blade leakage. The efficiency of each stage is less 
than that of the turbine as a whole since the increase in heat content 
due to friction, etc., is available for transformation into useful work in 
the succeeding stages. To find the actual pressure condition in each 
stage allowing for the various losses, it is necessary to correct the theo- 
retical quantities for these losses. See " Energy and Pressure Drop in 
Compound Steam Turbines," by F. E. Cardullo, Proc. A.S.M.E., Feb., 
1911, and paper read by Prof. C. H. Peabody, Proc. Society of Naval 
Architects and Marine Engineers, June, 1909. 

223. The Curtis Steam Turbine. — Figs. 252 to 265 show the general 
arrangement and a few details of the Curtis steam turbine, which is of 
the compound pressure and velocity type. The total expansion is 
carried out in one or more compartments or stages, each stage compris- 
ing a set of expanding nozzles and a wheel carrying two or more rows of 
buckets. A high initial velocity is given to the jet in each stage by 
expansion in the nozzles as in the De Laval, and the energy is absorbed 
by successive action upon the series of moving and stationary vanes 



STEAM TURBINES 



413 



arranged somewhat as in the Parsons turbines, paragraph 226. In the 
latter, however, the difference in pressure between the two sides of 
each vane induces flow by continuous expansion, while in the former 
the moving vanes in any one stage simply absorb the kinetic energy 
already created by expansion in the nozzle. The action is as follows: 
Steam enters stage (1), Fig. 252, through the first set of nozzles, and is 
partially expanded. With the resulting initial velocity it impinges 
against the first row of moving blades and gives up part of its energy, 
and is then deflected through the adjoining stationary blades to the 
next set of moving vanes, where its velocity is still further reduced, and 




J t|| 

Diaphragm T/^\ | 
[Wir^N and Second Afp^ *_-? ^ 
^ ^ Stage JSozzleJ^ ^ ^ 



Fig. 252. Arrangement of Nozzles and Blades, Curtis Turbine. 

so on until it has been brought practically to rest. From this stage 
the steam flows at reduced pressure through nozzles of second stage, 
which are sufficient in number and in size to afford the greater area re- 
quired by the increased volume. In expanding in these nozzles it 
acquires new velocity and gives up energy to the moving blades as 
before. This process is repeated through two to five stages, depending 
upon the size of turbine. Fig. 253 shows a partial section of a four- 
stage vertical 5000-kw. machine. R, R are sections through the revolv- 
ing wheels, which in this particular turbine are nine feet in diameter and 
keyed to the vertical shaft S. On the periphery of each wheel are 
bolted two rows of blades or vanes, with a stationary or intermediate 
row attached to the casing between them. The buckets are made of 
rolled nickel bronze, hammered to shape and finish. The roots are dove- 
tailed into the holders and the tips are tenoned and riveted into a 



414 



STEAM POWER PLANT ENGINEERING 




TOBSO-HLTERNATOR 




Fig. 253. Four-stage Vertical Curtis Turbo-Generator. Base Condenser Type. 



STEAM TURBINES 



415 



shroud ring, thus insuring positive spacing and a rigid construction. 
See Fig. 254. Between each pair of wheels is a stationary steam-tight 
diaphragm P, which contains the nozzles through which the steam is 
expanded from the preceding stage. It will be 
noticed that the buckets and nozzles increase 
rapidly in size in succeeding stages as the pres- 
sure falls and the volume of steam increases. 
The parts are so proportioned that the steam 

gives up approximately - of its energy in each 



stage, n representing the number of stages. 
The number of stages and the number of vanes 
in a stage are governed" by the degree of expan- 
sion, the peripheral velocity which is desirable 
or practicable, and by various conditions of 
mechanical expediency. The admission valves vary in number and 
in location with the size of turbine. The automatic stage valve G 
connects the first stage directly to a set of auxiliary second-stage 
nozzles. See Fig. 255. Thus the overload capacity is increased by 

widening the steam belt and 




Fig. 254. Bucket Details, 
Curtis Turbine. 




not by admitting high-pressure 
steam into an intermediate stage 
as was formerly the practice with 
Curtis turbines. This method 
of overload control results in 
higher efficiency than the older 
system. 

Curtis turbines appear to have 
a wider range of economical 
application than any other type, 
commercial sizes ranging from a 
small horizontal unit of 7 kilo- 
watts rated output to vertical 
units of 25,000 kilowatts capac- 
ity on the continuous 24-hour 
basis. The smaller machines, 
7500 kilowatts and under, are usually of the horizontal type, and the 
larger units, 9000 kilowatts and larger, are of the vertical type. All 
Curtis turbines are governed by " cutting-out nozzles"; that is, full 
initial pressure is maintained in all the nozzles that are open and 
the capacity of the machine is controlled by varying the number in 
operation. Units under 500 kilowatts are ordinarily controlled by a 



Fig. 255. 



Automatic Stage Valve, Curtis 
Turbine. 



416 



STEAM POWER PLANT ENGINEERING 



-i m 




STEAM TURBINES 



417 



mechanical valve gear and the larger units by an indirect or relay system. 
In the older types this relay system was electrically operated; in the 
modern machines the valves are hydraulically controlled. 

Fig. 257 shows a section through the governor for the large units. 
Speed regulation is accomplished by the balance maintained between 
the centrifugal force of moving weights AA and the static force exerted 
by spring D. The governor is provided with an auxiliary spring F for 
varying its speed when synchronizing, the tension in which is varied by 
a small pilot motor controlled from the switchboard. The movement 




Fig. 257. Main Governor, Curtis Turbine. 

of the governor weights is transmitted through rod C to arm H and by 
means of the latter to the controlling mechanism of the valve gear. 

Fig. 258 gives an assembly view of the mechanical valve gear as 
applied to a 300-kilowatt unit. The valve stems extend upward through 
ordinary stuffing boxes and are attached to notched crossheads 8, 8. 
Each crosshead is actuated by a pair of reciprocating pawls or dogs, 
6, 6, the lower one of which closes the valve and the upper one opens it. 
The several pairs of pawls are hung on a common shaft which receives 
a rocking motion from a crank driven by the turbine shaft. The cross- 
heads have notches milled in the side in which the pawls engage to 
open or close the valve, the engagement being determined by shield 



418 



STEAM POWER PLANT ENGINEERING 



plates 2, the positions of which are controlled by the governor through 
the medium of suitable levers. Shield plates 6 are set one a little ahead 
of the other to obtain successive opening or closing of the valves. The 
pawls are held in position when not in contact with the shield plates 
by springs W. 




Fig. 258. Assembly of Mechanical Valve Gears for 300-Kw. Curtis Steam Turbine. 

Fig. 259 gives a diagrammatic arrangement of the hydraulically 
controlled valve-gear mechanism. The motion of governor g is trans- 
mitted through lever i to lever a of the pilot valve j. Pilot valve j 
controls the supply of oil (under pressure) in cylinder k, the piston of 
which actuates rods I, I* The movement of rod I is transmitted through 

* The pressure is on both sides of the piston and the latter is actuated by re- 
leasing the pressure. 






STEAM TURBINES 



419 



rack m to a small pinion. This pinion is mounted on the end of a shaft 
fitted with a number of cams, one a little ahead of the other, each cam 
controlling the opening and closing of a steam valve through the medium 
of rocker arm /. As the load on the turbine increases the governor 
slows down and causes the cam shaft to rotate in a reverse direction, in- 





Turbtne 



Fig. 259. 



Diagrammatic Arrangement of Hydraulically Operated Valve Gear, Curtis 
Turbine. 



dicated by the arrow points in Fig. 259. This causes a proportionate 
number of valves to be lifted and held open, the number increasing as 
the load increases, until all are open. Should the load continue to in- 
crease, as in the case of overload, the secondary valve opens as previously 
described, connecting the first stage with a set of auxiliary second- 
stage nozzles. Only the nozzles in the first stage are controlled by the 



m 3 



12345678 

Steam Belt Area 

Fig. 260. Steam-belt Area in Five-stage Curtis Turbine. 

governor. Should the turbine run above normal speed the emergency 
stop-valve automatically closes the admission of steam to the nozzles. 
This device consists of a steel ring placed around the shaft between the 
turbine and the generator. This ring is eccentrically mounted and the 
unbalanced centrifugal force is balanced by a helical spring. When 
the predetermined speed is reached the centrifugal force overcomes the 



420 



STEAM POWER PLANT ENGINEERING 



Drain 




Fig. 



261. Step Bearing, Vertical 
Curtis Turbine. 



spring tension and the ring moves in a still more eccentric position. In 
this position the ring strikes a bell-crank lever which trips the throttle 
valve and permits it to close by its own weight and the unbalanced 

pressure on the valve stem. 

In the Curtis turbine the area of the 
steam admission is limited to a small 
portion of the circumference in the first 
stages and does not extend around the 
entire circumference until the last stage 
is reached. See Fig. 260. 

The step bearing of a vertical machine 
is illustrated in Fig. 261. The weight 
of the rotor is supported by oil under 
pressure forced between the bearing 
blocks F and P, thus permitting the 
shaft S to revolve on a film of oil. The 
smaller disk P is attached by dowels to 
the main shaft. Carbon packing rings 
are used above the bearing to prevent 
leakage, and adjustment is provided in 
the lower bearing block by means of 
worm gear D, worm G and ratchet handle C. The oil pressure varies 
from 150 to 750 pounds per square inch according to the size of 
machine, the higher pressures being used in the larger machines. The 
carbon packing and casing for the high-pressure end of the vertical 
turbine is shown in Fig. 262. Figs. 
263 and 264 give general details of 
the carbon packing and water- 
cooled bearings of the horizontal 
turbine. 

Fig. 265 gives a diagrammatic 
outline of the oiling system. A 
tank, of sufficient capacity to con- 
tain all the oil and fitted with suit- 
able straining devices and a cooling 
coil, is located at a level low enough 
to receive oil by gravity from all 
points lubricated. A pump draws 
oil from this tank and delivers it at a pressure about 25 per cent higher 
than that required to sustain the weight of the turbine in the step 
bearing. A spiral duct baffle connects the source of pressure to the 
step bearing and serves to regulate the oil supply to the lower end of 



ter Packing Casing 




Fig. 262. 



J^Shafft 

YTurbine Head 

Carbon Packing, Vertical Curtis 
Turbine. 



STEAM TURBINES 



421 



Carbon. Packing-^; 
Casing 

Drains 




Sealing Steam 



Fig. 263. Carbon Packing, Horizontal Curtis Turbine. 






Cooling Coil 



J LJ 
-i CD 


LJ 
CD 


^ 


LJ LJ 1 


cn^j r 


w 


cn 


CD CD 


□ □ L 








] a 


CD 


a ib 


°ftj[ 


I* 


en 


□ cj 


CD cd r 


J CD 


lZI 


CD CD 


□1&)) L 


N^ 


CD 


>|r a 


mEKL 


-1 a 


□ 


CDOCD 


CD 1 1) ) r 


2 m 

i i — i 






n J 


L_l 


=fe! 


CD d°b L 

i—ir-i r 


n 




OiLEeed Inlet 



Fig. 264. Water-cooled Bearing, Horizontal Curtis Turbine. 



422 



STEAM POWER PLANT ENGINEERING 



the shaft. This source of pressure is also connected through a reduc- 
ing valve to the upper oiling system of the machine, in which a pressure 
of about 60 pounds to the square inch is maintained. This system, 

which includes a storage tank 
partly filled with compressed air, 
operates the hydraulic governor 
mechanism and supplies oil to the 
upper bearings. Delivery of oil 
to these bearings is regulated by 
adjustable baffles designed to offer 
resistance to the oil flow without 
forcing the oil to pass through any 
very small opening which might 
easily become clogged. A relief 
valve is provided to prevent the 
pressure in the upper part of the 
oiling system from rising above a 
desirable limit. Drain pipes from 
the upper bearings and from the 
hydraulic cylinder and relief valve 
all discharge into a common cham- 
ber, in which the streams are 
visible, so that the oil distribution 
can always be easily observed. 
At some point in the high-pressure 
system adjacent to the pump it 



Relief Valve 




Air Chamber 

Gauge Glass 

For Charging 
Air Chamber 
Three Way Cock 
Check-Valve 
Reducing Valve 
Valve 



To Spring Equalizer 
or Accumulator. 



Storage 
Tank 



Fig. 265. Arrangement of Oiling System for is desirable to install a device to 

Curtis Turbine. equalize the delivery of oil from 

the pump, as is done by the air chamber commonly used with pumps 
designed for low pressure. A small spring accumulator is furnished for 
this purpose, except in cases where weighted storage accumulators are 
used. In large stations where several machines are installed, a storage 
accumulator is desirable and can be arranged advantageously so that 
it will normally remain full, but will discharge if pressure fails, and in 
doing so will start auxiliary pumping apparatus. 





DIRECT 


CURRENT. 




Kw. 


R.p.m. 


Kw. 


R.p.m. 


15 
25 
75 


4,000 
3,600 
2,400 


150 

300 
500 


2,000 
1,800 
1,500 



STEAM TURBINES 423 



ALTERNATING CURRENT. 



300 


1,800 


2,000 


900 




500 


1,800 


3,000 


) 




1,000 


1,200 


to 


600-750 




1,500 


900 


20,000 


J 





For the description of a -typical steam-turbine station equipped with 
Curtis turbines see Chapter XX. 

General Description of Curtis Turbines: Power, March, 1909; Engr. U. S., Jan. 1, 
1908, p. 115; Power & Engr., Feb. 25, 1908, p. 284, Feb. 25, 1908, March 3, 1908; 
Elec. Wld., June 17, 1905, p. 1136. 

Guide Bearings, Oil Distribution & Carbon Packing: Power & Engr., April 14, 1908; 
Jan. 3, 1911, p. 10. 

Mechanical Valve Gear: Power & Engr., March 10, 1908, p. 356. 

Hydraulic Valve Gear: Power, March, 1909, p. 189. 

Heat Losses: Power & Engr., Jan. 3, 1911, p. 19; Proc. A.S.M.E., Feb., 1911, p. 123. 

224t. Elementary Theory, Curtis Turbine. — Fig. 266 gives a dia- 
grammatic arrangement of the blades and nozzles in the first stage 
of a two-stage Curtis turbine, each stage consisting of one set of nozzles 
and two moving and one stationary sets of blades. 

Referring to the diagram: the steam is expanded in the first stage 
from pressure Pi to P 2 and issues from the first set of nozzles with absolute 
velocity Vi, striking the first set of moving blades at an angle a with 
the line of motion of the wheel. The resultant v± of Vi and the peripheral 
velocity u is the velocity of the steam relative to the vanes; and the 
angle jS which the line Vi makes with the line of motion of the wheel is 
the proper entrance angle of the blades for the first set. Neglecting 
friction the exit angle y will be the same as the entrance angle (3. The 
resultant of v 2 , the exit velocity relative to the blade, and u, the peripheral 
velocity, is V 2 , the absolute exit velocity. 

Since the second set of blades is fixed and serves as a means of chang- 
ing the direction of flow, the absolute velocity entering them is V 2 . 
The angle 8 formed by V 2 and the center line of the stationary blades 
is the proper entrance angle. Neglecting friction the absolute exit 
velocity will be V z = V 2 , and the exit angle will be e = 8. The steam 
flowing from the stationary blades strikes the second set of moving 
blades at an angle e = 8 with absolute velocity V s . Combining V 3 with 
the peripheral velocity u we get v 3 , the velocity of the steam relative to 
the second set of moving blades. The angle 6, formed by v 3 and the 
line of motion of the wheel, is the proper entrance angle for the second 
set of moving blades. The resultant of v± ( = v 3 ) and u is F 4 , the absolute 
exit velocity for the first stage. 



424 



STEAM POWER PLANT ENGINEERING 



In the second stage the steam is expanded from pressure P2 to that 
in the condenser and acquires initial velocity V aj leaving the last bucket 
with residual velocity V n . The theoretical velocities and blade angles 
for this stage may be found as above. 




Fig. 266. Velocity Diagram, Curtis Turbine. 

Example: A four-stage Curtis turbine develops 800 horse power 
on a steam consumption of 12 pounds per horse-power hour; initial 
pressure 150 pounds absolute, superheat 100 degrees F., back pressure 
1.5 pounds absolute, peripheral velocity 450 feet per second, angle of 
the nozzle with the plane of rotation, 20 degrees. Each stage consists 
of two rotating elements and one stationary element. Compare the 
performance of the actual turbine with its theoretical possibilities. 






STEAM TURBINES 425 

Ideal Turbine: 

For the sake of simplicity it will be assumed that the final velocity 
of each stage is zero and that the heat drop in the first set of nozzles is 
one fourth of the total theoretical drop assuming adiabatic expansion. 
From steam tables H 1 = 1249.6 B.t.u. 
From entropy tables or Mollier diagram H n = 934.6. 
Total heat drop = 1249.6 - 934.6 = 315. 
Heat drop in first stage ^ A = 78.75. 

The velocity of the jet in the first stage is 

V x = 224 V78.75 = 1985 feet per second. 

By laying off this initial velocity in direction and amount and com- 
bining it with the peripheral velocity as in Fig. 266, the absolute velocities 
F 2 and Vz may be readily obtained. 

The kinetic energy absorbed in the first set of moving blades, per 
pound of steam, is 

= tt^j; (1985 2 - 1170 2 ) = 39,930 foot pounds per second, 
and in the second set of moving blades 

^ = 6T4 (722 - F32) 

(1170 2 - 670 2 ) 



64.4 
= 14,280 foot-pounds per second. 

The total energy converted into useful work is 

39,930 + 14,280 = 54,200 foot-pounds per second. 

Had the entire heat drop been utilized in doing work the total energy 
would be 

■^t-j X 1985 2 = 61,180 foot-pounds per second. 

The difference 61,180 - 54,210 = 6970 represents the loss due to the 
residual velocity of the steam leaving the last bucket. 

Since the steam is brought to rest before entering the second set of 

£1070 

nozzles, the heat equivalent of this energy or = 8.96 B.t.u. in- 

77o 

creases the final heat content; thus 

# 2 = 1249.6 - 78.75 + 8.96 = 1179.8 B.t.u. 



426 STEAM POWER PLANT ENGINEERING 

But a total heat drop per stage of 78.75 B.t.u. was assumed as a 
requirement of the problem and the final result obtained above shows 
it to be 78.5 — 8.96 = 69.54. By trial and adjustment or by means of 
empirical formulas a value of H 2 may be obtained which will fulfil the 
given conditions. Such an analysis is beyond the scope of this book, 
and the reader is referred to Forrest E. Cardullo's article "Energy and 
Pressure Drops in Compound Steam Turbines," Trans. A.S.M.E., 
vol. 33, p. 325, 1911. 

The remaining stages may be analyzed in a similar manner. 

It should be borne in mind that in the actual turbine the velocity 
will be less than the theoretical on account of frictional resistances in 
the nozzles and blades and the heat content H 1} H 2 . . . H n will be 
greater than that of the ideal mechanism. Radiation, leakage, windage 
and other losses must also be considered in determining actual conditions. 

Neglecting the residual energy in the exhaust, the total heat drop 
Hi — H n is available for doing useful work and the water rate of the 
ideal turbine is 

w 2546 2546 Q1 , , , 

W = 77 yt = ~^tf =8.1 pounds per horse-power hour. 

Hi — ti n olo 

Heat consumption per horse-power hour per minute 
= 8.1(124^6 -83.9) = 157Rtu 

01) 

Thermal efficiency 

1249.6 - 934.6 _ 
r 1249.6-83.9 
Actual Turbine: 

Steam used per hour = 800 X 12 = 9600 pounds. 
Steam used per second = 9600 4- 3600 = 2.66 pounds. 
Horse power developed per pound of steam flowing per second = 
800 4- 2.66 = 300. 

Kinetic energy converted into useful work 

300 X 550 = 165,000 foot-pounds per second. 
Thermal efficiency, equation (138) 

2546 _ 

' 12 (1249.6 - 83.9) 

Heat consumption, B.t.u. per horse power per minute, 
12 (1249.6 - 83.9) _ 9QQ 
60 ~ " 2SS ' 

Efficiency ratio = y/ = * 07A = 0.675. 

£j r \).Z/U 



STEAM TURBINES 

SUMMARY 



427 



Horse power developed per pound of steam 

Steam consumption, pounds per horse-power hour 

B.t.u. consumed per horse power per minute 

Thermal efficiency, per cent 

Efficiency ratio, per cent 



Actual 
Turbine. 



300 

12.0 
233 

18.2 
67.5 



Perfect 
Turbine. 



445 

8.1 
157 
27.0 




Fig. 267. Section through Westinghouse Impulse Turbine. 

224. Westinghouse Impulse Turbine. — To meet the demand for 
small turbines the Westinghouse Machine Company has recently placed 
on the market turbines ranging in size from 25 to 500 horse power. 
Units under 50 horse power are of the single-stage multi-velocity type 
and the larger sizes are of the two-stage multi- velocity type. All of 
these turbines have but one moving element. Fig. 267 shows a section 
through one of the larger units and Fig. 268 a diagrammatic arrange- 
ment of blades and nozzles. Referring to Fig. 268, steam enters the 
turbine at A, through a double-seated poppet valve, and is partially 



428 



STEAM POWER PLANT ENGINEERING 



expanded in nozzle B. Leaving the nozzle it impinges on the moving 
blades, giving up part of its energy, and thence passes to the first re- 
versing chamber C, where its direction is changed and it is redirected 
against the moving blades, The steam is then expanded through the 
second set of nozzles to the existing back pressure and the cycle is 
repeated. Turbines of 50 horse power or less have but one set of 
nozzles and one reversing chamber, all others having two sets of nozzles 
and reversing chambers. 




Steam Inlet 
Fig. 268. Arrangement of Blades and Nozzles — Westinghouse Impulse Turbine. 

226. Westinghouse Single-flow Steam Turbine. — Fig. 269 shows a sec- 
tion through a Westinghouse turbine of the original Parsons design and 
illustrates the multi-stage reaction type. In this type no nozzles are 
employed and expansion of the steam is effected by a series of stationary 
and moving blades. The rotor is a steel barrel or drum divided into 
three sections of varying diameter, upon the periphery of which bronze 
blades are radially inserted in dovetailed grooves. The adoption of three 
sections of varying diameter has no bearing on the design of this machine 
but is merely for mechanical convenience. The blades increase in length 
and cross section from the high-pressure to the low-pressure end of each 
section. The stator is of cast iron and its inner surface is studded with 
rows of blades projecting radially inward and conforming in size with 
the adjoining blades of the rotor. The relative positions of the blades 
in the rotor and stator are shown in Fig. 270. The operation of the tur- 
bine is as follows: Steam enters at S, Fig. 269, through poppet valve 
V, which is actuated by the governor shown in detail in Fig. 273, 
and flows through the annular space between rotor and stator to the 
exhaust opening at B. The entire expansion is carried out within this 
annular compartment and resembles in effect a simple divergent nozzle 
with the exception that the dynamic relationship of jet and vane is such 
as to secure a comparatively low velocity from inlet to exhaust. The 



STEAM TURBINES 



429 




430 



STEAM POWER PLANT ENGINEERING 



velocity varies from 150 feet per second at the high-pressure end to about 
600 feet per second as a maximum at the low-pressure end. The action 
of the steam on the blades is illustrated in Fig. 270. The steam strikes 
the first set of stationary blades as at P with initial velocity of about 150 



Stationary 



CCUCC(C< 

)) )) )) ))';)) )) ro 



Blades 



Moving Blades 



z 



Vs 



Stationary 



ccucxuc . 

ZfflHHIH" 



Blades 



Blades 



Fig. 270. Flow of Steam in Parsons Turbine. 



feet per second and is deflected against the moving blades immediately 
adjoining. In passing from P to Pi the steam is partly expanded and 
gives up a portion of its energy to the moving blades. The steam is de- 
flected from Pi to Pn and thus has a reactive effect on the moving blades 
in addition to the impulse imparted at Pi. The total torque produced 



1 



■e 





Plan View 



Side View 



End. View 



Fig. 271. Old Method of Fastening Blades in Westinghouse-Parsons Turbines.* 

at the shaft in element A is therefore due to impulse from 1 and reaction 
from 2. This process is repeated in each element of the turbine, the steam 
expanding as it flows from element to element in its passage to the con- 
denser. Opposed to the three sets of blades the spindle also carries three 
rotating balance pistons P, P, Fig. 269, each of such diameter as to ex- 

* For present construction of interlocking root of blade see Jour. A.S.M.E., Aug., 
1912, p. 1157. 



STEAM TURBINES 



431 



actly balance, through passage E, the axial thrust of the steam against its 
corresponding drum of blades. 

Steam enters the turbine intermittently as shown in Fig. 272, which 
represents indicator cards from a 1250-kilowatt turbine at various loads. 



Atmospheric=0 
Absolute Zero 




Fig. 272. 



Indicator Cards Showing Initial Pressure in a Westinghouse-Parsons Steam 
Turbine. 



At light load the valve opens for a very short period. As the load in- 
creases, the period lengthens until finally, at about full load, the valve 
does not reach its seat at all, and continuous pressure is obtained in the 
high-pressure end of the turbine. 




Fig. 273. Governor Mechanism, Westinghouse-Parsons Turbine. 

The intermittent admission of steam is produced and controlled as 
follows: Lever T, Fig. 273, is given a reciprocating motion by an eccen- 
tric actuated by a worm and worm wheel on the main shaft. This 
motion is transmitted through lever H (with fixed fulcrum B) to lever A 



432 



STEAM POWER PLANT ENGINEERING 



(with floating fulcrum D) and finally to pilot valve G. This recipro- 
cating pilot valve admits puffs of steam from pipe to the under side of 
piston M, the rod R of which is attached to the admission valve V in 
Fig. 269. A spiral spring holds piston M in its lowest position until 
steam admitted by the pilot overcomes the spring tension and lifts the 
main valve from its seat, thereby permitting steam to enter the turbine. 
The fulcrum D of lever A is raised and lowered by the governor and 
therefore the pilot valve is controlled both by the motion of the eccen- 
tric and the motion of the governor. The eccentric keeps the pilot 
valve, and hence the main throttle, in constant oscillation, while the 
movement of the governor changes the limits of this motion. 



$w 




Fig. 274. Valve Gear with Steam Relay, Westinghouse-Parsons Turbine. 

If an overload is sufficiently great to cause the governor balls to drop 
to their lowest position, the auxiliary or secondary valve V a , Fig. 269, 
begins to open and admits high-pressure steam to the later stage where 
the working steam areas are greater, thus increasing in proportion the 
total power of the turbine. The operation of this valve is the same as 
the main admission valve and is controlled by the governor. Fig. 274 
shows the details of this mechanism. 

Fig. 275 shows the general details of the valve gear for the large units. 
Relay valve A is controlled by the governor and admits pressure oil to 
or exhausts it from the operating cylinders. When oil is admitted to 
the operating cylinder raising the piston, the lever C lifts the primary 
valve E. The lever D moves simultaneously with C, but on account 
of the slotted connection with the stem of the secondary valve F, the 
latter does not begin to lift until the primary valve is raised to the point 
at which its effective opening ceases to be increased by further upward 
travel, 



STEAM TURBINES 



433 



In the smaller-sized machines running above 1200 r.p.m. flexible bear- 
ings are employed to absorb the vibration incident to the critical velocity. 
They consist of a nest of loosely fitting concentric bronze sleeves with 
sufficient clearance between them to insure the formation of a film of 
oil. In the larger, machines a split self-aligning bearing is used instead 
of the flexible bearing. The ends of the casing are fitted with a water- 
seal made by a revolving wheel to prevent the escape of steam or inflow 
of air at the point of entry of the shaft. Leakage between dummy or 
balance cylinders is prevented by labyrinth packing which consists 
essentially of a series of grooves cut into the rotor into which bronze 
strips project from the casing. The steam in passing through the 




Fig. 275. Valve Gear with Oil Relay, Westinghouse-Parsons Turbine. 

packing is alternately wire-drawn and checked so that the flow is greatly 
reduced. 

221, Westinghouse Impulse-reaction or Double-flow Turbine. — In re- 
action turbines of the single-flow type, as illustrated in Fig. 269, the 
high-pressure portion dealing with the high-pressure incoming steam is 
the least efficient. This is due to the fact that the blade lengths are 
approximately proportional to the specific volume of the steam, and 
consequently the initial expansion in the turbine requires blade passages 
of very small dimensions. This results in greater leakage past the tips 
of the blades than in the low-pressure elements where the blades are 
long. Again, in the single-flow type the high-pressure balance piston 
occupies fully one half of the total balance-piston length of the shaft, 
while the low-pressure piston is 2\ times the high-pressure diameter 
so that balance pistons occupy a large portion of the total bulk of the 
machine. By making the high-pressure element of the impulse type 
and by arranging the low-pressure reaction elements on either side as 



434 



STEAM POWER PLANT ENGINEERING 




STEAM TURBINES 



435 



illustrated in Fig. 276 the efficiency may be increased and the 
the turbine may be greatly decreased. 
There are two rows of moving 
buckets upon the impulse wheel with 
an intermediate set of reversing 
blades, the operation being practi- 
cally the same as in the first stage of 
a Curtis turbine. The drop in pres- 
sure in the nozzles is such that 
approximately 20 per cent of 
the total energy developed is 
absorbed by this impulse ele- 
ment. After leaving the im- 
pulse element the steam 
divides, one portion passing 
directly to the low-pressure 
blading at the left, while the 
rest passes through the hol- 
low shell of the rotor to the tf 
similar pressure blades upon 
the right. As these sections 
are equal and symmetrical 
they counterbalance each 
other, and the balance or 
"dummy" pistons may 
be dispensed with. The 
advantages of the double-flow 
type over a single-flow unit of 
equal capacity are: (1) reduction 



bulk of 




nearly 50 per cent in the shaft span 
between bearings; (2) the diameters 
of the casing and rotating part 
are more uniform, thus tending to 
greater rigidity; (3) a reduction of 
about 70 per cent in the bulk of the 
main parts of the machine, and (4) 
internal stresses due to high-pressure 
and high-temperature steam are 
avoided by isolating the incoming steam, without separate nozzle 
chambers. 
228. Allis-Chalmers Steam Turbine. — Fig 



277 shows a 
through an Allis-Chalmers standard steam turbine, which is 



section 
of the 



436 



STEAM POWER PLANT ENGINEERING 



Parsons type but differs from the original Parsons machine and the 
Westinghouse-Parsons construction principally in manufacturing de- 
tails. In the older Parsons type, three balance pistons are placed at 
the high-pressure end. In the Allis-Chalmers design, the larger piston 
is placed at the low-pressure end of the rotor, behind the last row of 
blades, the other two remaining at the high-pressure end. This con- 
struction permits of a smaller balance piston and allows a smaller 
working clearance in the high-pressure and intermediate cylinders. In 
the Allis-Chalmers turbine the roots of the blades are dovetailed and 
fitted into a foundation ring, and the tips are incased in a channel- 
shaped shroud ring, thereby insuring a rigid and positively spaced con- 
struction. The governor is of the Parsons type, except that the main 



P=150 



Stationary 



Vanes M J J J 




Fig. 278. Velocity Diagram. Westinghouse-Parsons Turbine. 



valve and pilot valve are actuated by hydraulic instead of steam pres- 
sure. The bearings are of the self-adjusting ball-and-socket pattern 
and are kept "floating in oil" by a small pump geared to the turbine 
shaft. The oil is passed through a tubular cooler with water circula- 
tion after it leaves the bearings and is used over and over again. 

229. Elementary Theory, Parsons Turbine. — Fig. 278 gives a dia- 
grammatic arrangement of fixed and stationary blades in the first 
stages of a multi-stage ideal reaction turbine. The steam enters the 
stationary blades at practically zero velocity and is there partially 
expanded and impinges against the movable blades at velocity V h 
part of the energy of the steam being thus absorbed. In passing 
through the movable blades the steam is still further expanded and 
leaves at an absolute velocity F 2 , exerting an additional pressure on 
the blades from the reaction. The steam enters the second set of 
stationary blades with velocity V2 and is still further expanded to 
velocity Vz, and so on. 



STEAM TURBINES 437 

The energy imparted to the steam in the first set of stationary blades 
is 

El = ^g V ^ ( 157 ) 

Vi = absolute velocity of the steam leaving the blades. 
The energy imparted to the steam in the first set of moving blades is 

W 
E2=tt W ~ *>i 2 ). (158) 

Vi = relative velocity of the steam entering the moving blades. 
v 2 = relative velocity of the steam leaving the moving blades. 
The total energy acquired by the steam in the first stage is 

Ei + E 2 . 

The energy converted into work in this stage is 

WVo 2 
E = E X + E 2 - -V 1 - (159) 

= (V 1 *+v 2 *-v 1 *-V 2 *)^. (160) 

V 2 = absolute velocity of the steam leaving the moving blades. 

Each stage may be analyzed in a similar manner. 

Example: A Westinghouse-Parsons turbine develops 1000 horse 
power on a steam consumption of 12 pounds of steam per horse-power 
hour. Initial steam pressure 150 pounds per square inch absolute; back 
pressure 1 pound per square inch absolute; drop in pressure in each 
set of fixed and moving blades 15 pounds per square inch; peripheral 
velocity 300 feet per second; a x = a 2 = 30 degrees. Compare the 
performance of the actual and ideal turbine. 

Actual turbine: 

Steam consumed per hour, 

1000 X 12 = 12,000 pounds. 

Steam consumed per second, 

12,000 -i- 3600 = 3.33 pounds. 

Horse power developed per pound of steam flowing per second, 

1000 ^ 3.33 = 300. 

Kinetic energy per pound of steam, 

300 X 550 = 165,000 foot-pounds per second. 

Thermal efficiency, 

et " 2546 

^ = 12 (1193.4- 69.8) =18 - 8perCent ' 



= 66.9 per cent. 



438 STEAM POWER PLANT ENGINEERING 

Heat consumption, B.t.u. per horse-power hour per minute, 

12 (1193.4 - 69.8) _ 

60 " ^ 4 ' 7 ' 

Efficiency ratio, 

Et = 18.8 
E r 28.1 
Ideal turbine: 

The velocity imparted to the steam in the first set of stationary 
blades due to the drop from 150 to 135 pounds per square inch is 

V 1 = 224 VH 1 - H 2 

= 224 V1193.4 - 1184.8 
= 657 feet per second. 

Lay off the value of V\ in direction and amount and combine with u, 
the peripheral velocity, Fig. 278. The resultant is v h the velocity 
of the steam relative to the blades. The angle between v\ and the line 
of motion of the wheel will be the angle with the blade at entrance. 

From the velocity diagram, 

V! = 424. 

E 2 , the energy given up by one pound of steam in expanding from 
135 to 120 pounds, is 

E 2 = 778 (H 2 - H 3 ) 

= 778 (1184.8 - 1175.1) 

= 7554 foot pounds per second. 

Substitute V\ = 424 and E 2 = 7554 in equation (158), 

7554= 60 fe2 " 4242) ' 
v 2 =816 feet per second. 

The resultant of v 2 and u is V 2 , the absolute velocity of the steam 
leaving the moving blades of the first stage. From the diagram, 

V 2 = 573 feet per second. 

The energy converted into work in the first stage is determined by 
substituting the proper values in equation (160), thus: 

E = "(657 2 + 816 2 - 424 2 - 573 2 ) ^ 
= 9150 foot-pounds per second. 
The various stages may be analyzed in a similar manner. 



STEAM TURBINES 439 

The theoretical output of the entire turbine per pound of steam will 
be that corresponding to adiabatic expansion from a pressure of 150 to 

1 pound absolute. 

E = 778 (#i - H n ) 

= 778 (1193.4 - 877.2) 

= 246,000 foot-pounds per second. 

Horse power per pound of steam, 

H.P. = ^°00 = 444. 
550 

Steam consumption per horse-power hour, 

3600 



444 



8.1 pounds. 



Thermal efficiency, 

= 1193.4 - 877.2 

r ~ 1193.4 - 69.8 

= 28.1 per cent. 

The calculation of Parsons Turbines: Zeit. f. Turbine, May 10 and 20, 1912. 

230. Low-pressure Mixed-pressure and Bleeder Turbines. — A prom- 
ising field for the steam turbine is in its application as a secondary or 
low-pressure unit in connection with non-condensing or condensing 
engines, or, combined with a regenerator, in connection with engines 
using steam intermittently. Numerous examples may be cited showing 
great gains in both capacity and economy in existing power plarv+? 
involving the abandonment of but a negligible part of the equipment 
and accomplishing this result with a minimum additional investment. 
The most notable installation, to date, of low-pressure turbines to con- 
densing reciprocating engines is at the 59th Street Station of the Inter- 
borough Rapid Transit Co., New York. Three of the nine 7500-kw. 
Manhattan-type compound Corliss engines have been equipped with 
Curtis three-stage, low-pressure turbo-generators of equal capacity, and 
provision is made for the installation of six additional units. The low- 
pressure turbine is installed between the exhaust of the low-pressure 
cylinders and the condenser as shown in Fig. 279. Running with the 
engine the low-pressure turbine generator carries a variable load without 
governor regulation. The turbine generator takes care of the speed 
by automatically taking such a load as will keep the frequency in unison 
with that of the engine-driven generator. The turbine is equipped 
with the usual emergency speed-limit attachment for cutting off the 
steam supply should the speed exceed a predetermined limit. The 
performance of one set of engines, a high-pressure turbine of the equiva- 
lent total capacity, and that of the combined engine and low-pressure 



440 



STEAM POWER PLANT ENGINEERING 




iPSw. 279. Low-pressure Turbine Installation at the 59th Street Station of the Inter- 
borough Rapid Transit Company, New York. 




Fig. 280. Westinghouse Double-flow Low-pressure Turbine. 



STEAM TURBINES 



441 




Fig. 281. Performance of 7500-Kw. Engine at 59th Street Station of Interborough 
Rapid Transit Company, New York, with Varying Receiver Pressure. 

19 

















































Act 


aal 1 


'est) 


Eng 


ine 








































































__lGi 


Wai 


itee) 


mg 


r h Pressu 


re Turbi 


ne 












































































jt 














f* 




















ruit 


ine 


















B*~ 


Co 


nbin 


ed 1 


:ngu 


ie-ai 


fd-t 


)W~1 


>reat> 






































A:- Constant Nozzle Pressure 

@ 16 fr abs. 
B.:- Variable Nozzle Pressure 
All Nozzles Opeu 


























































Cor 
Tur 


rectec 
bine £ 


lfor 
>team 


Moist 


ure u 


i 



7000 



8000 



9000 



10000 11000 12000 13000 
Unit Load-X-W, 



14000 15000 16000 



Fig. 282. Comparison of Economy Curves: 7500-Kw. High-pressure Turbine, 7500-Kw. 
Engine and Combined Engine and Low-pressure Turbine at the 59th Street Station 
of the Interborough Rapid Transit Company, New York. 

turbine are illustrated in Fig. 282. The conclusions drawn from an 
exhaustive series of tests at this station are that the addition of low- 
pressure turbines effected: 

a. An increase of 100 per cent in maximum capacity of plant. 

b. An increase of 146 per cent in economic capacity of plant. 

c. A saving of approximately 85 per cent of the condensed steam for 
return to the boiler. 



442 



STEAM POWER PLANT ENGINEERING 



d. An average improvement in economy of 13 per cent over the best 
high-pressure turbine results. 

e. An average improvement in economy of 25 per cent (between the 
limits of 7000 kw. and 15,000 kw.) over the results obtained by the 
engine units alone. 

/. An average unit thermal efficiency of 20.6 per cent between the 
limits of 6500 kw. and 15,500 kw. 

If the turbine does not constantly require all the available exhaust 
from the high-pressure units the exhaust from the latter may be by- 
passed so that the high-pressure units may reap the benefit of the reduced 
back pressure when the turbine is carrying a light load. This bypass is 
controlled by the governing system. Similarly in case the exhaust from 




Fig. 283. Rateau Low-pressure Steam Turbine Installation. 

the high-pressure units is in excess of the turbine requirements the 
bypass valve discharges part of the steam directly to the condenser. 

Low-pressure turbines are frequently installed in connection with 
regenerative accumulators, to rolling-mill engines, steam hammers, and 
other appliances using steam intermittently, and have proved to be 
paying investments. A typical installation of this character is to be 
found at the South Chicago Division of the International Harvester 
Company. The front elevation of the turbine and regenerator installa- 
tion is shown in Fig. 283 and the general arrangement of the regenerator 
is shown in Fig. 284. The regenerative accumulator is intended to 
regulate the intermittent flow of steam before it passes to the turbine. 
The steam collects and is condensed as it enters the apparatus and is 
again vaporized during the time when the exhaust of the engines dimin- 
ishes or ceases. (See also Power & Engr., Nov. 8, 1910.) 



STEAM TURBINES 



443 



The regenerator consists of a cylindrical boiler-steel shell divided 
into two similar chambers by a central horizontal diaphragm. In each 
compartment are a number of elliptical tubes A, each of which is per- 
forated with a number of f-inch holes. The spaces surrounding the 
tubes and, under certain conditions, the tubes themselves are rilled 
with water to a height of about four inches above the top of the upper 
tubes. Baffle plate B serves to separate the entrained moisture from 
the steam. The operation is as follows: Exhaust steam enters the 
apparatus at N, passes to the interior of the elliptical tubes, and escapes 
into the steam space through the perforations and thence to the turbine. 
When the supply of steam from the main engine ceases, the pressure in 



Exhaust Sf earn "Inlet 1 




Fig. 284. Rateau Regenerator Accumulator. 



the regenerator decreases, the water liberates part of the heat it has 
absorbed and a uniform flow of low-pressure steam is given off. The 
continued demand of the turbine reduces the pressure in the accumula- 
tor and causes the steam still retained in the tubes to escape, thereby 
maintaining the circulation of the water (indicated by arrowheads) and 
facilitating the liberation of steam. Suitable valves regulate the limits 
of pressure in the accumulator and prevent the return of water to the 
main engine. 

In the size normally installed this type of accumulator will furnish 
a sufficient supply of steam for four minutes with exhaust entirely 
cut off. If the period is longer than four minutes it becomes necessary 
to admit live steam. Low-pressure turbines develop one electrical 
horse-power hour on a steam consumption of about 30 pounds with 
initial pressure of 15 pounds absolute and a back pressure of 1.5 pounds 



444 



STEAM POWER PLANT ENGINEERING 



absolute. Fig. 285 gives the performance of a typical Westinghouse 
low-pressure turbine for various vacua, initial pressure 15 pounds 
absolute. 

The weight of water W required to operate the low-pressure turbine 
for a given period with a predetermined temperature drop may be 
calculated from the relationship 

W = -^-., (161) 

Q\ -9% 



75 




I 




1500 B.H.P. 


Economy Test 
Westinghouse Low Pressure Turbine 




\ 




Water Rates per B.H.P. Hr. at different Vacua 
Speed 1800 R.P.M. 


70 


























\ 


























65 
W60 

a 55 

g50 

& 

i 40 

30 
35 




V 














«*•* 


^ 


-t ^?fit^— 








\ 


\ 






¥5 


^ 


e^ 
















V 


A 


























\ 


V 




2 


4" 


2 


j" 


2 
Full 


6" 
Load 


2 


r" 


2 


8" 






\\ 








V 


acuun 


i m in 


chest C 


0'Bai 


o.) 
















^ 


*" ^ 






























Wsj 


^Per 


£#.J 


l^2i- 




























__26* 




























_28" 








20 





































100* 
90 # 
80# 



200 400 600 800 1000 1200 1400 1600 1800 2000 2200 2400 2600 
Load - Brake Horse power 

Fig. 285. Performance of Westinghouse Low-pressure Turbine. 

in which 

t = maximum number of minutes the exhaust supply may be entirely 

cut off; 
s = water rate of the turbine, pounds per minute; 
r = mean latent heat at regenerator pressure; 
qi = heat of the liquid corresponding to maximum temperature of 

water in regenerator, degrees F.; 
q 2 = heat of the liquid corresponding to minimum temperature of 
water in regenerator, degrees F. 

If the regenerator is to absorb M pounds of exhaust steam in t minutes 
as in case of a sudden flux of exhaust the weight of water Wi required is 

Mr 



Wi = 



Qi -q* 



(162) 



STEAM TURBINES 



445 



Example: Determine the weight of water to be stored in a regenerator 
to operate a 500-horse-power exhaust steam turbine for five minutes 
if the steam supply is entirely cut off; pressure drop 17 to 14 pounds 
absolute, turbine water rate 30 pounds per horse-power hour. 

Here 

• t = 5j 8 = 500X30 = 250| r = 965.6 + 971.9 _ ^ 



60 

qi = 187.5, q 2 = 177.5, 



W = 



5 X 250 X 968.8 
187.5 - 177.5 



= 121,100. 



If the regenerator is to absorb 2000 pounds of the exhaust steam in five 
minutes during a period of sudden flux, 

2000 X 968.8 



Wi = 



187.5 - 177.5 



= 193,760. 




W* 



Fig. 286. Westinghouse Mixed-pressure Turbine. 

Low-pressure turbines equipped with special expanding nozzles, or 
the equivalent, to receive steam at high pressure direct from the boilers 
are known as mixed-pressure turbines. With this construction the full 
power of the turbine can be developed with (1) all low-pressure steam, 
(2) all high-pressure steam, (3) any proportion of high- and low-pressure 
steam. In the Curtis mixed-pressure turbine this transition from all 
low pressure to all high pressure, through all the conditions intermediate 
between these extremes, is provided for automatically by the turbine 
governor; a deficiency of low-pressure steam causes the high-pressure 
nozzles to open automatically. With this arrangement it is not neces- 
sary for purposes of economy to proportion exactly the low-pressure 



446 STEAM POWER PLANT ENGINEERING 

turbine to the amount of exhaust steam available, but within limits it 
may be made as large as the load demands. 

Fig. 286 shows the general details of a Westinghouse mixed-pressure 
turbine. 

Mixed-pressure turbines have been constructed in single units as 
large as 7000 kw. (Gen. Elec. Rev., July, 1912.) 

In compound condensing engine plants it has been general practice 
to draw steam from the receiver for heating and manufacturing pur- 
poses. This same result is effected in the bleeder type of turbine, through 
the agency of a pressure-controlled valve placed between the high- and 
low-pressure section of the turbine and which automatically diverts the 
required amount of steam at a predetermined pressure to the heating 
S3 r stem. 

Mixed Pressure Turbines and Engines: Power, Feb. 1, 1910; Eng. Rec, Feb. 19, 
1910; Eng. Mag., Apr., 1909; National Engr., May, 1912. 
Geared Turbines: Engng., May 24, 1912. 
Heating in Connection with Steam Turbines: Power, Sept. 17, 1912, p. 426. 

231. Advantages of the Steam Turbine. — The principal advantages 
of the steam turbine are: (1) simplicity; (2) economy of space and 
foundation; (3) absence of oil in condensed steam; (4) freedom from 
vibration; (5) uniform angular velocity; and (6) high efficiencies for 
large variations in load. The reciprocating engine is well adapted for 
pumping stations, direct-current generators, compressor plants, hoist- 
ing engines, and the like, requiring low angular velocity, but its place 
is being rapidly taken by the steam turbine for alternating-current 
dynamos, centrifugal pumps and blowers requiring high angular ve- 
locity. The recent development of high-efficiency speed-reduction gear- 
ing makes it possible for the turbines to compete with the engines for 
low-speed work. 

Simplicity. — Although composed of a large number of parts as com- 
pared with a reciprocating engine of the same capacity, there are few 
moving parts and rubbing surfaces. The only contact between rotor 
and stator is in the main bearings, and the problem of lubrication is 
therefore a simple one. The absence of pistons, stuffing boxes, dash 
pots, etc., reduces the cost of maintenance and attendance to a minimum 
and limits the possibility of leakage. 

Economy of Space and Foundation. — The floor space required by 
practically all types of turbines is considerably less than the space 
requirements of piston engines. Vertical three-cylinder compound 
Corliss engines of the New York Edison type require the least floor 
space of any large slow-speed reciprocating engines, but take up about 
twice the space of a Parsons turbine installation of the same size. With 



STEAM TURBINES 447 

high-speed engines of the Willans central-valve type the comparative 
economy in space is less marked. 

The weight of the steam turbine is very small compared with a re- 
ciprocating engine of the same horse power. The New York Edison 
engine and generators weigh more than eight times as much as a turbine 
installation of equal capacity. The turbine, for this reason, and also 
because of the total absence of vibration, requires a relatively light 
foundation. In many instances the foundation consists of steel beams 
with concrete arches sprung between them resting upon the floor, and 
the basement underneath may be used for the condenser instead of the 
massive foundation required for the reciprocating engine. Engines are 
seldom constructed in sizes above 10,000 horse power, whereas single 
turbine units of 20,000 kw. are not uncommon and a Parsons turbine 
of 25,000 kw. normal capacity is now being installed in the Fisk Street 
Station of the Commonwealth Edison Co., Chicago. 

Absence of Oil in Condensed Steam. — As the steam turbine requires 
no internal lubrication, oil does not come in contact with the steam, 
and the condensed steam from the surface condensers is available for 
boiler-feeding purposes without purification. In many cases the re- 
use of condensed steam effects a large saving in cost of feed water and 
in expense for maintenance and cleaning of boilers. The amount of 
entrained air is reduced to a minimum and consequently the work of 
ai-r pumps is lessened. 

Regulation. — The variable pressure at the crank pin of a recipro- 
cating engine necessitates the use of a heavy flywheel to keep the in- 
stantaneous angular fluctuation within practical limits. In the steam 
turbine the motion is purely rotary and a flywheel is not necessary. 
In the former there are always instantaneous variations in velocity 
during each revolution, even with constant load, while in the latter the 
speed is practically constant. A number of published tests of Par- 
sons and Curtis turbines show an average fluctuation of 2 per cent from 
no load to full load and 3 per cent from no load to 100-per-cent over- 
load. Although closer regulation than this is possible, it is not deemed 
necessary, particularly in alternating-current work where a compara- 
tively wide range is desirable for parallel operation. 

Overload Capacity. — The overload capacity of any prime mover de- 
pends entirely upon the designation of the rated load. The maximum 
economy of the average piston engine lies between 0.7 and full load, 
and for this reason the rated load refers usually to this maximum eco- 
nomical load. Evidently if the engine is rated under its maximum pos- 
sible output it is capable of overload. Under the existing system of rating 
the average piston engine is capable of operating with overloads of 



448 STEAM POWER PLANT ENGINEERING 

25 to 50 per cent. According to the old rating the steam turbine was 
capable of overloads ranging from 100 to 200 per cent and much con- 
fusion arose in determining the station load factor. Current turbine 
practice gives as the normal rating the maximum continuous load 
which can be carried for 24 hours when under control of the primary 
valves. Through the agency of the secondary valves overloads of 50 
per cent or more are possible. The steam economy of the turbine is 
superior to that of the engine for overloads. 

232. Efficiency and Economy of Steam Turbines. — As far as steam 
consumption is concerned there is practically no difference between 
the performance of standard high-grade piston engines and that of 
first-class steam turbines (both using saturated steam) for sizes under 
2000 kilowatts, the choice depending more upon variable load character- 
istics and space requirements than upon heat economy. Engines of 
the uniflow type are more economical in steam consumption than tur- 
bines of equivalent capacity and piston engines using highly super- 
heated steam are decidedly more economical of fuel than turbines 
under the best conditions of operation, but heat economy is only one 
of the items entering into the ultimate cost of power. In a general 
sense the turbine gives a flatter load characteristic with saturated steam 
than the standard piston engine * and for this reason is better adapted 
to variable loads, but this advantage disappears with the use of highly 
superheated steam. For sizes over 2000 kilowatts the turbine is in a 
class of its own and piston engines above this size are seldom found in 
modern stationary practice. A comparison of Fig. 198 showing t}^p- 
ical economy curves of high-speed single-valve non-condensing engines, 
and of Fig. 289 showing similar curves for small non-condensing tur- 
bines is somewhat in favor of the piston engine, though the difference 
is small; whereas a comparison of the turbine and engine curves in 
Fig. 282, showing the performance of very large units, is decidedly in 
favor of the turbine. Any number of individual tests may be cited 
showing superiority in fuel consumption of the piston engine over that 
of a turbine of equivalent capacity and vice versa, but when the ma- 
chines are designed for the same operating conditions the results are 
practically the same for all sizes under 2000 kilowatts. Tables 58 to 74 
give the general condition of operation and the steam consumption of 
exceptionally good piston engines of various sizes and types, and Table 
75 similar data of first-class turbines. A study of these tables will show 
that the choice must be based on other factors than the steam con- 
sumption. In a general sense, the piston engine is superior to the tur- 
bine for high back pressures, slow rotative speeds and heavy starting 
torques, while the turbine has practically superseded the engine for 
* This applies to very large units only. 



STEAM TURBINES 



449 






TtdONIMHOOOOiOOONOOLOOiCO^OWiOaiO^CiMOO 



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450 



STEAM POWER PLANT ENGINEERING 



large central station units and for auxiliaries requiring high rotative 
speed. Recent tests of the Melville reduction gear (Machinery, Feb., 
1910) show exceptionally high efficiencies for sizes as large as 6000 



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50 GO 70 

Brake Horse Power 



90 



Fig. 287. Typical Performance of a 90-horse-power Terry Steam Turbine. 

kilowatts, and it is not unlikely that the turbine equipped with this 
device will offset the low rotative speed factor of the piston engine. 

If the tests of steam turbines and piston engines could be made at 
some standard initial pressure, back pressure and quality or superheat, 



























































































































































































































































































1 
























































































































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G000 



7000 



8000 9000 
Load in K.W. 



10000 



11000 



12000 



Fig. 288. Typical Performance of 9000-kilowatt Curtis Turbine; 200 pounds Gauge 
Pressure, 125 degrees Superheat, 29 Inches Vacuum. 

then a comparison could readily be made, but both types of prime 
movers are designed to give the best results for special operating con- 
ditions, and any marked departure from these conditions will result 
in loss of economy. It is frequently desired, however, to make a 



STEAM TURBINES 



451 



comparison between the economy of the different machines, and the 
following methods are in vogue : 

(1) Steam consumption under assumed conditions. 

(2) Heat consumption per unit output per minute above the ideal 
feed-water temperature. 

(3) Efficiency ratio or ratio of actual to ideal. 



70 

u 

« 60 



>0 

§50 

& 

II 



40 



■30 



20 



Steam Press. -150 Lb. 
Dry Steam 
Atmospheric Exhaust 




Load 
Fig. 289. Economy Tests of Small Non-condensing Turbines. 

Standard Correction Curves: 

This method for comparing engines or turbines or both is best illus- 
trated by a specific example: Suppose it is required to compare the 
full-load performance of a 125-kilowatt direct-connected piston engine 
with that of a 125-kilowatt turbo-generator with operating conditions 
as follows : 





Steam Consump- 
tion, Lbs. per 
Kw.-Hour. 


Initial Pres- 
sure, Lbs. 
Absolute. 


Vacuum, 

Inches of 

Hg. 


Superheat, 
Deg. F. 


Engine 


25.0 

22.7 


160 
110 


25.5 

28.0 





Turbine 


125 







Manufacturers of steam turbines have provided correction curves as 
illustrated in Fig. 290, showing the influence of varying vacuum, super- 
heat and pressures on the steam consumption.* From curve B, we 

* These curves are drawn to a much larger scale than the reproduction given here. 



452 STEAM POWER PLANT ENGINEERING 

find that the steam consumption of the turbine should be decreased 
2.5 pounds to give the equivalent at 160 pounds initial pressure; from 
curve A it should be increased 2.5 pounds to give the equivalent at 
25.5 inches of vacuum, and from curve C it should be increased 2.5 
pounds to give the equivalent at degree superheat. The full-load 
steam consumption for the turbine under the engine conditions is 
therefore 22.7 - 2.5 + 2.5 + 2.5 = 25.2 pounds per kilowatt-hour. 

The ratio method is also used in this connection, thus: The full-load 
steam consumption at 160 pounds pressure, curve B, Fig. 290, is multi- 

25 

plied by the ratio -^r-= to give the equivalent consumption at 110 pounds 

(25 is the steam consumption at 160 pounds and 27.5 the consumption 

at 110 pounds). Similarly the correction ratio to change the con- 

25.5 
sumption at 28 inches of vacuum to 25.5 is -^f~, and to correct 125 

25 
degrees F. superheat to degree F. is~~-=- 

Summary. 

25 
Pressure correction ==-r = 0.91 = — 9 per cent. 

Z i .Q 

27.5 
Vacuum correction -^r- = 1.10 = 10 per cent. 

25 

Superheat correction ^-r = 1.11 = 11 per cent. 

Net correction 12 per cent. 

Corrected steam consumption = 22.7 + 0.12 X 22.7 = 25.4 pounds per 
kilowatt-hour. 

The ratio method is generally used if the difference between the 
corrected steam consumption and that of the correction curves for 
the same conditions is greater than 5 per cent ("The Steam Turbine," 
Moyer, p. 128). 

This ratio method for correcting steam consumption at full load may 
be used without appreciable error for half to one and one half load and 
is the only practical method for quarter load. (Engineering, London, 
March 2, 1906.) 

Heat Consumption: 

The heat consumption B.t.u. per unit output per minute above the 
ideal feed-water temperature may be expressed 

W(Hi-q 2 ) ^ See equation (136). 
oO 



STEAM TURBINES 



453 



Steam Pressure* L"bs. per Sq. In. Absolute 
120 130 140 150 160 170 



200 




100 120 

Superheat, Deg. Fan. 
23 24 25 

Vacuum, Inches of Mercury 

Fig. 290. 



























































































s 


140,000 




















































































s 












©- Guarantee points 
A - Water rate-low vacuum as tested 
B'- "Water rate-high vacuum as tested 
C- Water rate-corrected to 175 lb. pressure, 
100 deg. superheat, 28 in. vacuum 




































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18 

17 S 

15 a3 



1,000 2,000 3,000 4,000 5,000 6,000 7,000 8,000 9,000 10,000 11,000 

Load jn Kw. 

Fig. 291. Performance of 10,000-Kilowatt Westinghouse Double-flow Turbine, City 
Electric Co., San Francisco, Cal. 



454 



STEAM POWER PLANT ENGINEERING 



For the case cited above 

25 (1194.1 



Engine, 



Turbine, 



98) 



60 
22.7 (1264.2 - 70) 



60 



= 455 B.t.u. 
= 451 B.t.u. 



Efficiency Ratio: 

The efficiency ratio, or the extent to which the theoretical possibilities 
are realized, may be expressed 
2546 X 1.34 



Er W (Hi -H 2 ) 
For the case cited above 



See equation (139). 



Engine, 
Turbine, 



2546 X 1.34 
25 (1194.1 - 915) 
2546 X 1.34 



= 0.49. 



0.43. 



22.7 (1264.2 - 915.3) 

In the assumed case the turbine is the more economical in heat 
consumption, but the engine is the more perfect of the two as far as 
theoretical possibilities are concerned. 

A Recent Comparison of Turbines and Engines: Eng. Rec, Feb. 12, 1911; Power, 
Feb. 1, 1910. 

Comparing Steam Turbine Tests: Power, May 2, 1911. 

Testing Steam Turbo-Generators: Proc. A.I.E.E., Dec, 1910. 

Test of a 10, 000-Kilowatt Turbo-Generator Set: Jour. A.S.M.E., Dec, 1910. 

The Present State of Development of Large Steam Turbines: Jour. A.S.M.E., May, 1912. 

233. First Cost. — Steam turbines, generally speaking, are about 10 per 
cent lower in first cost than high-grade compound engines of equivalent 
power though the price depends largely upon the type and conditions 
for which they are designed. The figures in Table 76 refer to the cost of 
the high-pressure turbines direct connected to generators exclusive of 
auxiliaries and give some idea of the unit cost for various sizes. 

TABLE 76. 

APPROXIMATE COST OF STEAM TURBINES AND GENERATORS. 
In Dollars per Kilowatt. Rated Capacity. 





Kilowatts. 




25 

55 
60 


50 

47 
51 


75 

42 
46 


100 

38 
43 


200 

32 
36 


300 

32 
36 

35 
35 


500 


1000 


2000 


4000 


6000 


Direct current: 
Nori-condensing .... 








Condensing 












Alternating current: 
25 cycles 


32 
30 


28 
27 


25 
25 


21 
21 


20 


60 cycles 












20 

















STEAM TURBINES 



455 



234. Cost of Operation. — Data pertaining to first cost of operating 
steam-turbine and reciprocating-engine plants and combinations of both 
will be found in Chapter XVIII. The following table, contributed 
by H. G. Stott, Superintendent Motive Power of the Interborough 
Rapid Transit Company, New York, gives an excellent comparison of 
the relative maintenance and operating costs (to date) of the three 
types of steam power plants as applied to large central stations for 
electric street railways. 

TABLE 77. 

RELATIVE COSTS PER KILOWATT-HOUR. DISTRIBUTION OF 
MAINTENANCE AND OPERATION. 





Reciprocating 
Steam 
Plant. 


Steam 

Turbine 

Plant. 


Reciprocating 
Engines and Low- 
pressure Steam 
Turbines. 


Maintenance. 
1. Engine room, mechanical 


2.59 
4.65 
0.58 
1.13 

61.70 
7.20 
6.75 
7.20 
2.28 
1.07 
2.54 
1.78 
0.30 
0.17 


0.51 
4.33 
0.54 
1.13 

55.53 
0.65 
1.36 
6.74 
2.13 
0.95 
2.54 
0.35 
0.30 
0.17 


1.55 


2. Boiler or producer room . 


3.55 


3. Coal and ash handling apparatus. . . . 

4. Electrical apparatus 


0.44 
1.13 


Operation. 
5. Coal 


46.48 


6. Water 


0.61 


7. Engine room labor 


4.06 


8. Boiler or producer room labor 

9. Coal and ash handling labor 

10. Ash removal 


5.50 
1.75 
0.81 


11. Electrical labor. . 


2.54 


12. Engine room lubrication 


1.02 


13. Engine room waste, etc 


0.30 


14. Boiler room lubrication, etc 


0.17 






Relative operating cost, per cent. . . . 

Relative investment, per cent 

Probable average cost per kw 

Probable fixed charges 


100.00 
100.00 
125.00 
11% 


77.23 

75.00 
93.75 
11% 


69.91 

80.00 

100.00 

11% 





For steam-turbine plants larger than 60,000 kilowatt the cost per kilo- 
watt may be reduced to $65.00. 

235. Influence of Superheat. — The use of superheated steam in- 
creases the economy of the reciprocating engine about 1 per cent for 
every 10 to 20 degrees of superheat, depending upon the conditions of 
operation, the gain being due mainly to the reduction of cylinder con- 
densation. Cylinder condensation is reduced not only because of the 
excess heat available for the evaporation of moisture but also because 
superheated steam has a lower conductivity than wet steam, and less 
heat is given up to the cylinder walls for the same difference of tem- 
perature. In the steam turbine this difference of temperature is much 
smaller, since high- and low-pressure steam do not alternately come in 



456 STEAM POWER PLANT ENGINEERING 

contact with the same surface as is the case with the reciprocating 
engine, and the time of contact is considerably less, due to the com- 
paratively high velocities. With a well-lagged casing, therefore, the 
condensation due to this cause is insignificant compared with that of 
the reciprocating engine, and the beneficial effect of superheat is much 
more pronounced. Friction of the steam, which in the reciprocating 
engine is negligible, and which may be a source of considerable loss in 
the turbine, is greatly reduced by the use of superheated steam, as is 
also the " windage" loss due to the rapid revolution of the wheels. 

The problem of cylinder lubrication is sometimes a difficult one in 
steam engines using a high degree of superheat, and trouble is frequently 
experienced due to the unequal expansion of the metal. In the steam 
turbine the latter difficulty is not so pronounced and no internal lubrica- 
tion is necessary, hence a high degree of superheat is permissible. 
For maximum economy the steam at the end of expansion should be 
free from moisture. Assuming purely adiabatic expansion, the steam 
in expanding from 165 pounds to 1 pound absolute would have to be 
superheated about 1300 degrees F., giving the steam an actual tempera- 
ture of 1800 degrees F. A study of some 100 tests made in this country 
gives about 250 degrees superheat as a maximum and 100 degrees to 
150 degrees F. as an average. In Europe reciprocating engines are 
operating with superheat as high as 450 degrees F. and turbines 300 
degrees F.* The additional fixed and operating costs of superheating 
must be considered in determining the net gain, since the decrease in 
steam consumption does not represent the actual saving. With pres- 
sures of 175 pounds gauge or less, and not to exceed 200 degrees F. 
superheat, the net gain has in most cases proved a substantial one. 
With higher temperatures and pressures the cost of maintaining the 
superheat may increase more rapidly than the saving in steam consump- 
tion, until a limit is reached beyond which no advantage is gained. 
This is illustrated in Fig. 292. (From paper read by E. D. Dreyfus 
before the Railway Club of Pittsburgh, May 20, 1910.) 

236. Influence of High Vacua. — The possible economy of the recip- 
rocating engine is greatly restricted by its limited range of expansion. 
Cylinders cannot be profitably designed to accommodate the rapid 
increase in the volume of steam when expanded to very low pressures. 
For example, the specific volume of 1 pound of steam under a vacuum 
of 29 inches (referred to a 30-inch barometer) is about 667 cubic feet, 
or nearly double its volume under a vacuum of 28 inches. Usually 
the exhaust is opened at a pressure of 6 or 8 pounds absolute and con- 

* For results of European turbine tests with various degrees of superheat see 
Power & Engr., Feb. 14, 1911, p. 288. 






STEAM TURBINES 



457 



sequently a large proportion of the available energy is lost. The lower 
vacuum in the exhaust pipe, therefore, serves only to diminish the back 
pressure and does not affect the completeness of expansion. Even if it 
were practical to expand to 1 pound absolute, the increased condensa- 
tion in the reciprocating engine would offset any gain due to expansion 
unless the steam were highly superheated. A study of a number of 
tests of reciprocating engines shows a slight improvement due to in- 
creasing the vacuum beyond 26 inches. Tests of steam turbines show 



15 

I 








































^y 


















P. 

X 

I 10 










^ 


Spr 








Full Load Operation 
!75 o lDS.Steam Press 








< 


f' 












Cp Values from | 

Knoblauch & Jakob 


*5 




4 


'J/T 










>14. 


iljG 1 

Investment &.Main 
Charge Probated As 


;enan< 
sumin 
il-Exp 


1 


; 


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incr 


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easing 


LOUS -138 

Radia 


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sses 


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wltl 


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6 


8 


3 1C 


IS 


14 


t0 16 


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22 












Superheat-Deg-.Fahr. 
Fig. 292. Influence of Superheat on Overall Economy of Operation. 

a decrease of 3 to 4 per cent in steam consumption for each inch increase 
of vacuum between 25 and 29 inches, for with a well-lagged casing 
cylinder condensation is practically absent, since the high- and low- 
temperature steam do not alternately come in contact with the metallic 
surfaces as is the case with the reciprocating engine. 

Since the volume of the steam increases very rapidly with the decrease 
in back pressure the corresponding capacity and power required by the 
air and circulating pumps becomes proportionately larger. There is 
consequently a point where the improvement in steam economy fails 
to exceed the increased power demanded by the auxiliaries. This is 



458 



STEAM POWER PLANT ENGINEERING 



illustrated in Fig. 293 taken from the discussion by E. D. Dreyfus on 
a paper entitled "Test of a 10,000-kw. Steam Turbine," S. L. Naphtaly, 
Proc. A.S.M.E., Dec, 1910. The following notes refer to Fig. 293: 

200-kilowatt turbine using surface condensers supplied with average 
injection water at 55 degrees F. Steam conditions 175 pounds, 100 
degrees superheat. Most economic arrangement of auxiliaries selected 
for each vacuum to give best heat balance. Relation between invest- 
ment and fuel economy pro-rated on the assumption of the fixed charges 
being one half of fuel expense burning coal at $3.00 per ton. 

18 

* 16 

© 

•—i 

'§ a 12 



10 
























/ 




















/ 

i 
t 

/ 


















V 


/ 


















/ 
/ 

/ 


















/ 
/ 


n 










Difference due to increased 
auxiliary consumption ^- 


/ 






















D 
















































><< 


7 








c 7 















































35 26 27 28 29 

Vacuum at turbine exhaust 
30 inch barometer 

Fig. 293. Influence of Vacuum on Power. 

A = Actual reduction at the turbine. 

B = Net reduction in plant fuel consumption. 

C = Equated cost (greater fixed charges and maintenance) of obtain- 
ing the higher vacuum. 

D = Final plant improvement. 

237. Tesla Bladeless Turbine. — Fig. 294 shows a section through 
a 200-horse-power experimental turbine designed by Nikola Tesla for 
which extravagant claims have been made. It consists of a rotor com- 
posed of 25 steel disks (each £% inch thick and arranged on the shaft 
so that the length of the shaft covered by the disks is approximately 
3.5 inches) revolving in a plain cylindrical casing. There are no guide 



STEAM TURBINES 



459 



plates or vanes and the viscosity and adhesion of the steam is depended 
upon for driving the rotor instead of impulse and reaction as in the 
standard type of turbine. Steam flows from the circumference to the 
center, and, when the rotor is at rest, flows by a short curved path, as 
indicated by the line in the end view, across the face of the disk. When 
the rotor is up to speed the steam passes to the exhaust in a spiral path 
from 12 to 16 feet in length. Since the direction of rotation is deter- 
mined solely by the direction of the entering jet it is only necessary to 
change the direction of the latter to effect complete reversal of the rotor. 
Mr. Tesla states that a 200-horse-power turbine of this type has attained 
a performance of 38 pounds per horse-power hour, initial pressure 
125 pounds gauge, atmospheric exhaust, 9000 r.p.m. (Prac. Engineer, 
U. S., Dec, 1911, p. 852.) The space occupied by this unit is only 




Fig. 294. Tesla Bladeless Turbine. 

2 feet by 3 feet and 2 feet high and the weight of the engine alone is 
2 pounds per horse power developed. If the development of the future 
bears out Mr. Tesla's prediction this type of prime mover will revo- 
lutionize the steam-turbine industry. At this writing (Nov. 1912), 
however, very little information is available concerning the present 
status of this apparatus. 

238. "Spiro" Turbine. — Fig. 295 gives the general details of a new 
type of steam motor which is a sort of compromise between the rotary 
engine and the steam turbine. It consists essentially of a pair of 
herringbone gears revolving in a twin cylindrical casing. Steam enters 
space a, Fig. 295, through ports pp and presses upon the gear teeth,, 
driving them forward. The volume is increased from that indicated 
at a to that shown at b, c, d, e, and / and the energy produced is the 
product of the pressure and volume. Exhaust occurs when the ends 
of the grooves in which the action lies pass the line of contact so that 



460 



STEAM POWER PLANT ENGINEERING 



they are no longer closed by the teeth of the opposite gear. Rotation 
is effected by both pressure and impulse, although no attempt is made 
to produce a considerable pressure drop between the steam chest and 
the admission pressure. The load may be varied by throttling or by 
cutting off the steam supply. The "Spiro" is built in various sizes 
ranging from 1 to 200 horse power and operates at 2000 to 3000 r.p.m. 
The following tests give an idea of the economy effected by this type 
of motor. (Power, Feb. 6, 1912, p. 188.) 



Test 1. 



Test 2. 



Boiler pressure, pounds gauge 

Inlet pressure, pounds gauge 

Back pressure, pounds gauge 

Horse power developed 

R.p.m 

Steam, pounds per horse-power hour 



120 


130 


101.5 


115 


Atmos. 


Atmos 


25.3 


151 


2450 


2710 


53.2 


31.8 




Pig. 295. The " Spiro!! Turbine. 






CHAPTER XL 

CONDENSERS. 

239. Genetal. — A pound of dry steam at atmospheric pressure (29.92 
inches mercury) occupies a volume of 26.79 cubic feet. Suppose 
these 26.79 cubic feet of steam were contained in a closed vessel, and 
that the steam was subsequently condensed and its temperature lowered 
by suitable means to, say, 110 degrees F. The condensed steam would 
occupy only about T7 \o of its original volume, and the pressure would 
fall to 2.59 inches of mercury, the latter pressure being due to the ten- 
sion of the aqueous vapor at the given temperature. That is to say, 
the best vacuum theoretically attainable under the given conditions 
would be 29.92 — 2.59 = 27.33 inches. The lower the temperature to 
which the condensed steam is reduced the more nearly perfect will be 
the vacuum attained. 

If air is mixed with the steam the vacuum will be still more imperfect. 
Thus, suppose the vessel to contain one pound of steam and one-tenth of 
a pound of air under atmospheric pressure. The volume of the closed 
vessel in this case must be 26.79 + 1.69 = 28.48 cubic feet. (1.69 
= volume of T V pound of dry air at 212 degrees F. and 29.92 inches of 
mercury pressure.) 

After the steam has been condensed and its temperature reduced to 
110 degrees F. the absolute pressure in the condenser will be 4.1 inches, 
thus: 

According to Dalton's Laws: (1) The mass of a given kind of vapor 
required to saturate a given space at a given temperature is the same 
whether the vapor is all by itself or associated with vaporless gases; 

(2) the maximum tension of a given kind of vapor at a given temperature 
is the same whether it is all by itself or associated with vaporless gases; 

(3) in a mixture of gas and vapor the total pressure is equal to the sum 
of the partial pressures. The final pressure P c in the vessel is therefore 
the combined pressure of the air P a and that of the water vapor P v , or, 
assuming complete saturation, 

Pc=P a + P v . (163) 

and the final volume of the entrained air V a will be that of the vessel, 
which in the specific case under consideration is V a = V c = 28.48 
cubic feet. 

461 



462 STEAM POWER PLANT ENGINEERING 

The final pressure of the air in the vessel may be calculated from the 
well-known physical law 

TT-TT (164) 

in which 

Pi and P a = absolute pressures corresponding to absolute tem- 
peratures T 1 and T a ; 
Vi and V a = volumes corresponding to absolute temperatures 
7\ and T a . 

Here Pi = 29.92 V l = 1.69 T 1 = 460 + 212 = 672, 

V a = 28.48 T a = 460 + 110 = 570. 

Substitute these values in equation (164) 

29.92 X 1.69 = P a X 28.48 
672 570 

from which P a = 1.51. 

The final pressure in the vessel is, equation (163), 

P c = 1.51 + 2.59 =4.1 inches of mercury. 

Example: If the absolute pressure in a condenser is 4 inches of mer- 
cury and the temperature of the air-vapor mixture is 100 degrees F., 
required the percentage of air by weight in the mixture. 

From steam tables the pressure of vapor corresponding to a tem- 
perature of 100 degrees F. is 1.93 inches of mercury. 

Hence, from equation (163), 

P c =P a + P v , 
4 =P a + 1.93, 
P a = 2.07. 

Let V = volume of the condenser chamber, cubic feet. 
Then 0.00285 V = weight of vapor in the chamber (0.00285 = den- 
sity of water vapor at 100 degrees F.), and 

0.08635 X |^ X ^Q^Qo v = 0-°0491 V = weight of dry air in the 

chamber. (0.08635 = density of air at degrees F. and 29.92 inches of 
mercury pressure.) 

The total weight of the mixture is 

0.00285 V + 0.00491 V = 0.00776 V, 
and the percentage of air in the mixture is 

0.00491 V A „ 00 ao 
0.00776 V = °' 632 ° r 63 ' 2 Per Cent - 






CONDENSERS 463 

In practice air is always present in exhaust steam. A condenser is a 
device in which the process of condensation and subsequent removal of 
the air and condensed steam is continuous, the degree of vacuum obtained 
depending upon the tightness of valves and joints, the quantity of en- 
trained air, and the temperature to which the condensed steam is reduced. 

The 'degree of vacuum may be expressed in different ways. (1) Ex- 
cess of the atmospheric pressure over the observed vacuum. For 
example, a 26-inch vacuum implies that the pressure of the atmosphere 
is 26 inches of mercury above the pressure in the condenser. (2) Per 
cent of vacuum, by which is meant the ratio of the observed vacuum to 
the atmospheric pressure. Thus, with the barometer standing at 30 

26 
inches, a vacuum of 26 inches may be expressed as 100 X ™ = 86.6 

per cent vacuum. This method of expression gives an idea of the 
efficiency of the condensing system. For example, the degree of 
vacuum indicated by 26 inches would be 93 per cent with a barometric 
pressure of 28 inches but only 84 per cent when the barometer reads 
31 inches. (3) Absolute pressure. Thus a 26-inch vacuum referred to 
a 30-inch barometer would be indicated as a pressure of 30 — 26 = 4 
inches absolute, or 1.99 pounds per square inch. 

The mean atmospheric pressure at sea level is 14.7 pounds per square 
inch, corresponding to a mercury column 29.92 inches in height, tem- 
perature of the mercury 32 degrees F. If the reading of the vacuum 
gauge and of the barometer are both corrected to 32 degrees F., the 
difference gives the absolute pressure in inches of mercury. This is 
the usual method in average scientific investigations. In condenser 
practice it is customary to refer the readings of the vacuum gauge to a 
30-inch barometer in which case it is necessary to increase the standard 
temperature of the mercury to such a figure as will increase the height 
of the mean barometer from 29.92 to 30 inches, viz., 58.4 degrees F. 
Thus, if the barometer and the vacuum gauge readings are corrected 
to a temperature of 58.4 degrees F., the difference between the figures 
will give the absolute pressure in inches of mercury at 58.4 degrees F., 
and if the figure is subtracted from 30 inches, it will give the inches of 
vacuum referred to a standard barometer of 30 inches. 

The mercury column correction for any change in temperature may 
be closely approximated by the equation 

. , . , h = h l [l -0.00010101 -0L 

in which L -" 

h = height of mercury column corrected to temperature t; 

hi = observed height of mercury column; 

h = observed temperature of mercury column; 

t = temperature to which column is to be referred. 



464 STEAM POWER PLANT ENGINEERING 

Example: Height of mercury in vacuum gauge 28.52 inches, tem- 
perature of mercury 80 degrees F., barometer 29.85 inches, temperature 
42 degrees F. ; determine the vacuum referred to a 30-inch barometer. 
For the vacuum gauge 

h = 28.52 [1 - 0.000101 (80 - 58.4)] 
= 28.46. 
For the barometer 

h = 29.85 [1 - 0.000101 (42 - 58.4)] 
= 29.9. 
Absolute pressure in inches of mercury at temperature 58.4 degrees F. 
= 29.9 - 28.46 = 1.44. 
Vacuum referred to 30-inch barometer = 30 — 1.44 = 28.56. 

Properties of Air and Steam Mixtures in Relationto Condensing Plant: Engng., Jan. 
19, 1912. 

The Influence of Air on Vacuum in Surface Condensers: Engng., Apr. 17, 1908; 
Power, Feb. 2, 1909, p. 235; Nov. 22, 1910; Jour. A. S. M. E., Nov., 1912. 

Properties of Dry, Saturated and Unsaturated Air: Jour. Frank. Inst., Feb., 1911. 

240. Function of the Condenser. — The function of a condenser in 
connection with a steam engine or turbine is primarily the reduction of 
back pressure, though, in some instances, notably in marine work, the 
recovery of the condensed steam may be of equal importance. The 
advantages to be gained by decreasing back pressure may be most 
readily illustrated by the following example: A non-condensing engine 
taking steam at a pressure of 100 pounds absolute and cutting off at 
one-quarter stroke will have, theoretically, a mean effective pressure 
on the piston of 44.6 pounds per square inch, the back pressure being 
14.7 pounds per square inch absolute. If the engine exhausts into a 
condenser against a 26.5-inch vacuum (1.7 pounds absolute) the mean 
effective pressure will be increased to 44.6 + (14.7 — 1.7) = 57.6 
pounds per square inch, resulting in a gain in power which may be 
expressed 

P AS 
in which H - R= 3P00' (165) 

H.P. = horse power gained; 

P r = reduction in back pressure, pounds per square inch; 
A = area of the piston in square inches; 
S = piston speed in feet per minute. 

If P = mean effective pressure on the piston when running non-con- 
densing, the percentage of increase of power may be expressed 

Percent =100^- (166) 



CONDENSERS 



465 



In the above example the percentage of power gained would be 

1Q 

100 77-77 = 29.2 per cent. 
44.6 

The actual gain due to the use of the condenser would be much 

less than this, depending upon the type of engine and conditions of 

operation, as shown in the results of engine performances outlined in 

Chapter X. 

TABLE 78. 

PRESSURE OF AQUEOUS VAPOR IN INCHES OF MERCURY FOR EACH DEGREE F. 

(Marks and Davis.) 



30° 

40° 

50° 

60° 

70° 

80° 

90° 

100° 

110° 

120° 

130° 

140° 



o° 



248 
362 
522 
739 
03 
42 
93 
60 
44 
4.52 
5.88 



.257 
.376 
.541 
.764 
1.06 
1.46 
1.98 
2.66 
3.53 
4.64 
6.03 



.180 
.268 
.390 
.560 
.790 
1.10 
1.51 
2.04 
2.74 
3.63 
4.76 
6.18 



3° 



.188 
.278 
.405 
.580 
.817 
1.13 
1.55 
2.11 



2.82 
3.74 
4.89 
6.34 



.195 

.289 

.420 

.601 

.845 

.17 

.60 

.17 

.90 

.84 

.02 



6.51 



.203 
.300 
.436 
.622 
.873 
1.21 
1.65 
2.24 
2.99 
3.95 
5.16 
6.67 



c° 



.212 
.312 
.452 
.644 
.903 
1.25 
1.71 
2.30 
3.07 
4.06 
5.29 
6.84 



.220 
.324 
.468 
.667 
.964 
1.30 
1.76 
2.37 
3.16 
4.17 
5.43 
7.02 



.229 
.336 
.486 
.690 
.996 
1.33 
1.81 
2.44 
3.25 
4.28 
5.58 
7.20 



.238 
.349 
.503" 
.714 
1.03 
1.37 
1.87 
2.51 
3.34 
4.40 
5.73 
7.38 



With steam turbines the advantage gained by reduction of back 
pressure is more marked than with the reciprocating engine, though 
theoretically the same for the same range of expansion. Initial con- 
densation, leakage past valves, and other sources of loss prevent a 
reciprocating engine from benefiting from a good vacuum to the same 
extent as a turbine. 

Referring again to the example given above, if the steam is cut off at 
about one-sixth stroke, the work done when running condensing will be 
the same as when running non-condensing and cutting off at one-quarter. 
Theoretically the steam consumption will be decreased nearly in pro- 
portion to the reduction in cut-off. Generally speaking, a condensing 
engine will require from 20 to 30 per cent less steam for a given power 
than a non-condensing engine. (See results of engine tests, paragraph 
206.) This decrease in steam consumption is only an apparent one. 
If steam is used by the auxiliaries in creating the vacuum, the amount 
must be added to that consumed by the engine, unless the steam ex- 
hausted by the former is utilized to warm the feed water, in which case 
only the difference between the heat entering the auxiliaries and that 
returned to the heater should be charged against the engine. The power 



466 



STEAM POWER PLANT ENGINEERING 



necessary to operate the condenser auxiliaries varies from one to six 
per cent of the main engine power, depending upon the type and con- 
ditions of operation. (See paragraph 260.) 

In power plants where the exhaust steam is not used for heating or 
manufacturing purposes, the engines are almost invariably operated 
condensing, provided there is an abundant supply of cooling water. 
Even if the water supply is limited, it is often found to be economical 
to use some artificial cooling device, notwithstanding the high first cost 
and cost of operation of the latter. 

Some of the considerations affecting the propriety of running con- 
densing and the choice of condensing systems are taken up in para- 
graphs 263 and 264. 

241. Classification of Condensers. — The following is a classification 
of a few well-known condensers: 



r Standard low 



1. Jet condensers 



Parallel current (a) 



_ Counter current (6) . . 



Water cooled (a) 



( Worthington. 
\ Blake. 



2. Surface condensers 



Vacuum X Deane 

^hon teer*- 

Sector |f£*£ 

f Barometric j Kger- 

\ i Leblanc. 

[ High vacuum . . . < Wheeler. 

( Worthington. 

( Single-flow Baragwanath. 

< Double-flow Wheeler. 

( Multi-flow Wainwright. 

", A,-- „™i^ fh\ 5 Forced draft Fouche. 

| Air cooled (6) j Natupal draft penneU 

L Evaporative (c) Ledward. 



Condensers may be divided into two general groups: 

1. Jet condensers, in which the steam and cooling water mingle and 
the steam is condensed by direct contact, Figs. 296 to 303. 

2. Surface condensers, in which the steam and cooling medium are 
in separate chambers and the heat is abstracted from the steam by con- 
duction, Figs. 308 to 310. 

Jet condensers may be further grouped into two classes, according to 
the direction of flow of the air and cooling water: 

(a) Parallel-current condensers, in which the condensed steam, cool- 
ing water, and air flow in the same direction, collect at the bottom of 
the condenser chamber, and are exhausted by a suitable pump, Fig. 296. 

(b) Counter-current condensers, in which the cooling water and con- 
densed steam flow from the bottom of the chamber, while the air is 
drawn off at the top, Fig. 302. 



CONDENSERS 467 

Parallel-current condensers may be subdivided into three classes: 

(1) Standard condensers, in which the cooling water, condensed 
steam, and air are exhausted by a vacuum pump, Fig. 296. 

(2) Siphon condensers, in which the cooling water, condensed steam, 
and air are exhausted by a barometric column, Fig. 299. 

(3)' Ejector condensers, in which the condensed steam and air are 
exhausted by the cooling water on the ejector principle, Fig. 300. 

Surface condensers may be classified according to the nature of the 
cooling medium as 

(a) Water-cooled condensers. 

(b) Air-cooled condensers. 

(c) Evaporative condensers, in which the condensation of the steam is 
brought about by the evaporation of a fine stream of water trickling on 
the outside of the tubes. 

242. Standard Low-vacuum Jet Condensers. — Fig. 296 shows a 
section through a Worthington jet condenser, illustrating the parallel- 
current principle. When the pump is started a partial vacuum is 
created in the suction chamber above the valves H, H in the cone F. 
As soon as sufficient air has been exhausted, cooling water enters at B 
with a velocity depending upon the degree of vacuum in chamber F 
and the suction head, and is divided into a fine spray by the adjustable 
serrated cone D. The spray mingles with the exhaust steam entering 
at A and both move downwards with diverse velocities. The steam 
gives up its heat to the water and condenses. The velocity of the steam 
diminishes in its downward path to zero, while the velocity of the water 
increases according to the laws of falling bodies. The condensed steam, 
cooling water, and air collect at the lower part of the condenser and are 
exhausted by the wet air pump G, from which they are forced through 
opening J to the hot well. The vacuum in chamber F will depend upon 
the vapor tension of the warm water in the bottom of the well, the 
amount of air carried along by the cooling water and steam, and the 
tightness of valves and joints. In case the water accumulates in 
the condenser cone F, either by reason of an increased supply or by a 
sluggishness or even stoppage of the pump, the condensing surface is 
reduced to a minimum, as soon as the level of the water reaches the 
spray pipe and the spray becomes submerged, and only a small annular 
surface of water is exposed to the exhaust steam. The vacuum is 
immediately broken, and the exhaust steam escapes by blowing through 
the injection pipe and through the valves of the pump and out the dis- 
charge pipe at J, forcing the water ahead of it; consequently flooding of 
the steam cylinder cannot occur. In starting up the condenser a partial 



468 



STEAM POWER PLANT ENGINEERING 



vacuum for inducing a flow of injection water into the condenser cham- 
ber may be created by the pump if the suction lift is not too great. 
Many engineers, however, prefer to install a small forced injection or 
priming pipe the function of which is to condense sufficient steam to 
produce the necessary partial vacuum. 




Fig. 296. Worthington Independent Jet Condenser. 



Fig. 297 shows a section through the condensing chamber and air 
pump of a Blake vertical jet condenser with an automatic vacuum- 
breaking device. The injection water enters at opening marked " injec- 
tion" and flows through the adjustable " spray" nozzle in a fine spray, 
at an angle of about 45 degrees, and impinges on the conical sides of the 
upper condenser chamber. The spray falls from the sides to the pro- 
jecting ledges shown in the illustration. The ledges prevent the spray 



CONDENSERS 



469 



from falling directly to the bottom of the chamber and insure an efficient 
mingling of steam and cooling water. A perforated copper plate is 
substituted for the shelves when the force of the injection water is not 
sufficient to produce spray. The circulating water and condensed 
steam together with the non -condensable gases are drawn off at the 
bottom of the chamber. The vacuum-breaking device is shown at the 
right of the figure. When the rising water reaches the level of the float 



Steam 
Cylinder 



Hand Wheel ' 




Air Pump 
Fig. 297. Section through a Blake Jet Condenser. 

chamber, as in the case of an accidental stoppage of the air pumps, the 
float is raised and forces a check valve from its seat and allows an inrush 
of air to break the vacuum, thus preventing further suction of water 
into the condenser and consequent flooding of the engine. A is the forced 
injection or "priming" inlet used in starting up when the suction lift is 
considerable. 

Condensation of Steam: Cassier's Mag., May, 1912; June, 1911. 
Jet versus Surface Condensers: Power, March 21, 1911, p. 443; Engr., Lond., 
Dec. 23, 1910. 

Condensers: Power, March 19, 1912. 



470 STEAM POWER PLANT ENGINEERING 

243. Condensing Water, Jet Condensers. — In a jet condenser the 
cooling water and exhaust steam mingle, and the degree of vacuum is 
a function of the final or discharge temperature; thus the quantity of 
cooling water required depends upon its initial temperature, the tem- 
perature of the discharge water, and the total heat in the steam entering 
the condenser. If the steam in the low-pressure cylinder at exhaust is 
dry and saturated, the heat entering the condenser will correspond to 
the total heat in steam at exhaust pressure, but it usually contains con- 
siderable moisture, part of which is reevaporated when the exhaust 
valve opens to the condenser; however, it is sufficiently accurate for 
most practical purposes to assume the exhaust steam entering the con- 
denser to be dry and saturated and its heat to correspond to the pressure 
in the condenser. 

Let H = heat content of the steam at condenser pressure, 
t 2 = temperature of the discharge water, 
t = initial temperature of the cooling water, 
W = weight of cooling water in pounds necessary to condense 
and cool one pound of steam to the required discharge 
temperature. 

Then W = H ~ k t 32 • * (167) 

Example: How many pounds of cooling water are necessary to con- 
dense one pound of steam under the following conditions: Barometer 
29.92; vacuum 26 inches; temperature of injection water 60 degrees F. 

The temperature of aqueous vapor corresponding to an absolute 
pressure of 29.92 — 26 = 3.92 inches of mercury is 125 degrees F. (See 
Table 78.) The discharge temperature, however, must be less than 
this, as the pressure in the condenser is due not only to the aqueous 
vapor but to that of the air carried over with the circulating water and 
the condensed steam. In a condenser of the standard low-vacuum 
type the discharge temperature will be from 10 degrees to 15 degrees 
lower than that corresponding to the vacuum as recorded by the gauge. 
In this case assume it to be 15 degrees lower, i.e., t 2 = 125 — 15 = 110 
degrees. 

The total heat corresponding to a pressure of 3.92 inches of mercury 
is 1114 B.t.u. above 32 degrees (see steam tables); t = 60 degrees; 

t 2 = 110 degrees. 

= 1H4-110 + 32 = 
110-60 

* This expression is not theoretically correct, since it assumes a constant value 
of unity for the mean specific heat of water. The variation, however, is so slight 
that it may be neglected for all practical purposes. 



CONDENSERS 



471 



Evidently the higher the temperature of the discharge water the less 
will be the quantity of cooling water required, and consequently the 
smaller the weight of air introduced into the condenser; but the warmer 
the discharge water the greater will be the vapor tension and the lower 



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Fig. 298. 



60 70 80 90 

Temperature Degrees Fahrenheit 

Curves showing Relation Between Cooling Water and Hot-well Temperatures. 



the degree of vacuum. For reciprocating engines a hot-well tempera- 
ture between 110 degrees and 130 degrees F. is average practice; with 
turbines the temperature ranges between 80 degrees and 100 degrees F. 
On account of the inefficient heat absorption in practical installations, 
from 5 per cent to 15 per cent is added to the theoretical weight of cool- 
ing water as determined from equation (167). With jet condensers of 
the Leblanc type the temperature of the hot well is approximately 



TABLE 79. 

RATIO, BY WEIGHT, OF COOLING WATER TO STEAM CONDENSED (THEORETICAL). 











(Barometer 29.92.) 












Vacuum 24''. 


: 


Vacuum 25". 


Temp. 


Temperature of Steam 141°. 




Temperature of Steam 134°. 




Temp. 




of In- 
jection. 


Temperature of Hot Well. 


of In- 
jection. 


Temperature of Hot Well. 




HO 
15.0 


115 


120 


125 


130 


105 


110 


115 


120 


125 


40 


13.9 


12.9 


12.1 


11.4 


40 


16.1 


14.9 


13.8 


12.9 


12.1 


50 


17.5 


16.0 


14.8 


13.7 


12.8 


50 


19.0 


17.4 


16.0 


14.8 


13.7 


60 


21.0 


18.9 


17.3 


15.8 


14.6 


60 


23.2 


20.9 


18.9 


17.2 


15.8 


70 


26.2 


23.2 


20.7 


18.7 


17.1 


70 


30.0 


26.1 


23.0 


20.7 


18.7 


80 


35.0 


29.8 


25.9 


23.0 


20.5 


80 


42.0 


34.8 


29.6 


25.9 


22 8 


90 


52.4 


49.7 


34.6 


29.5 


25.6 


90 


70.0 


52.1 


41.5 


34.5 


29.4 





Vacuum 26 // . 
Temperature of Steam 125°. 


Temp, 
of In- 
jection. 


Vacuum 27". 
Temperature of Steam 114°. 


Temp, 
of In- 
jection. 


Temperature of Hot Well. 


Temperature of Hot Well. 




100 


105 


110 


115 




90 


95 


100 


105 




40 
50 
60 
70 
80 


17.5 
21.0 
26.3 
35.0 
57.6 


16.1 
19.0 
23.2 
30.0 
42.0 


14.8 
17.4 
20.9 
26.0 
34.7 


13.8 
16.0 
18.8 
23.0 
29.6 




40 
50 
60 
70 
80 


21.2 
26.5 
35.3 
52.9 


19.1 
23.4 
30.1 
42.1 


17.4 
20.9 
26.2 
34.9 
52.3 


16.0 
19.0 
23.2 
29.8 
41.5 













Temp, 
of In- 
jection. 



40 
50 
60 
70 



Vacuum 27.5". 
Temperature of Steam 108 c 



Temperature of Hot Well. 



80 


85 


90 


95 


26.6 


23.6 


21.1 


19.1 


35.6 


30.3 


26.4 


23.4 


52.3 


42.5 


35.2 


30.0 




70.8 


52.8 


42.0 



Temp, 
of In- 
jection. 



40 
50 
60 

70 



Vacuum 28". 
Temperature of Steam 100 c 



Temperature of Hot Well. 



75 80 85 90 



30.5 
42.7 
71.2 



26.6 
35.5 
53.2 



23.5 
30.2 
42.3 
70.6 



21.1 
26.3 
35.1 

52.7 



Temp, 
of In- 
jection. 


Vacuum 28.5". 
Temperature of Steam 90 


o 


Temp. 
of In- 
jection. 


Vacuum 29". 
Temperature of Steam 77 


3 


Temperature of Hot Well. 


Temperature of Hot Well. 




60 


65 


70 


75 




55 


60 


65 


67 




35 
40 
45 
50 


42.2 
54.0 
72.0 


35.8 
43.0 
53.5 
72.0 


30.6 
35.6 

42.8 
53.5 


29.2 
33.4 
38.8 
46.6 




35 
40 
45 
50 


52.0 
69.3 


43.0 
54.0 
71.5 


35.8 
43.0 
54.0 
72.0 


33.4 

38.4 
47.0 
61.0 













CONDENSERS 473 

5 degrees lower than that corresponding to the degree of vacuum in the 
condenser for inlet temperatures above 50 degrees F. For lower tem- 
peratures of inlet water the hot-well temperature ranges from 10 to 20 
degrees below that corresponding to the degree of vacuum. Table 79 
and the curves in Fig. 298 have been calculated from equation (167). 

244. Effect of Aqueous Vapor upon the Degree of Vacuum. — The 
futility of attempting to better the vacuum by exhausting the vapor is 
best illustrated by a specific problem. 

Required the volume of aqueous vapor to be withdrawn per hour 
from a condenser operating under the following conditions, in order 
that the vacuum may be increased one pound per square inch: Tem- 
perature of discharge water 125 degrees; corresponding vapor tension 
4 inches of mercury; barometer 30 inches; relative vacuum 26 inches; 
horse power 100; steam consumption 20 pounds per horse-power hour; 
cooling water 25 pounds per pound of steam condensed. 

100 X 20 X 25 = 50,000 pounds of cooling water per hour. 
= 833 pounds of cooling water per minute. 

Now to increase the vacuum one pound per square inch, approxi- 
mately 2 inches of mercury, the temperature of the water must be 
lowered to 102 degrees F., that is, 833 (125 - 102) = 19,159 B.t.u. 

19 159 
must be abstracted from the water in one minute, or ' =18.6 

pounds of water to be evaporated per minute. (1030 = average heat 
of vaporization of water under 26 to 28 inches of vacuum.) Now, one 
pound of vapor at 102 to 125 degrees F. has an average volume of 270 
cubic feet. 

Therefore 18.6 X 270 = 5022 cubic feet of vapor must be exhausted 
per minute to increase the vacuum from 26 to 28 inches, which is mani- 
festly impracticable. 

245. Injection Orifice. — The velocity of water entering a jet con- 
denser, neglecting friction, may be determined from the formula 



-here 



V = V2gh, (168) 



V = velocity of the water in feet per second, 
g = acceleration of gravity = 32.2, 
h = total head in feet. 

If p = pressure below the atmosphere in pounds per square inch, 

hi = distance in feet between the source of supply and the 
injection orifice, 

then h = 2.3 p ± h h (169) 



474 STEAM POWER PLANT ENGINEERING 

and equation (168) may be written 

V = 8.025 V2.3 p ± K (170) 

If the supply is under pressure, hi is positive; if under suction, it is 
negative. 

Example: What is the theoretical velocity of water entering a con- 
denser with 26-inch vacuum (referred to 30-inch barometer); suction 
head 8 feet. 

Here p = pressure in pounds per square inch, corresponding'' to 26 
inches of mercury = 12.8 pounds per square inch. 

h = 8. 

V = 8.025 V2.3 X 12.8 - 8 
= 37.1 feet per second 
= 2226 feet per minute. 

In proportioning the injection orifice in practice the maximum 
velocity of flow is assumed to be between 1500 and 1800 feet per minute, 
or, approximately, area of injection orifice in square inches = weight of 
injection water in pounds -=- 650 to 780. (" Manual of Marine Engineer- 
ing," Seaton, p. 204.) A rough rule gives area of orifice = area of low- 
pressure piston in square inches -5- 250. (Seaton, p. 204.) 

246. Volume of the Condenser Chamber. — According to Thurston 
the volume of a jet condenser should be from one fourth to one half 
that of the low-pressure engine cylinder. (" Steam Engine Manual," 
Thurston, II, 127.) 

According to Hutton the volume should not be less than that of the 
air pump and should approximate three fourths of that of the engine 
cylinder in communication with it. 

247. Injection and Discharge Pipes. — In practice the diameter of 
the injection pipe is based on a velocity of 400 to 600 feet per minute 
and that of the discharge pipe of 200 to 400 feet per minute; the lower 
figures for pipes under 8 inches in diameter, the upper range for larger 
diameters. 

(Atmospheric relief valves. — See paragraph 39.) 

248. Siphon Condensers. — Fig. 299 shows a section through a 
Baragwanath siphon condenser, illustrating the principles of a parallel- 
current barometric condenser. The cooling water enters the side of 
the condenser chamber at A and passes downward in a thin annular 
sheet around the hollow cone D. The exhaust steam enters at B and 
is given a downward direction by the goose neck C. It flows through 
the nozzle D and is condensed within the hollow cone of moving water, 
the combined mass including the entrained air discharging through the 
contracted throat E at high velocity into the tail pipe F. The water 



CONDENSERS 



475 



column in the tail pipe must be enough to overcome the pressure of the 
atmosphere; i.e., it should be 34 feet or more above the surface of the 
hot well, otherwise water would rise within this pipe to a height cor- 
responding to that of the barometer, which is approximately 34 feet for 
a barometric pressure of 30 inches of mercury. This is not strictly true 
when the condenser is in full operation, as the injector effect of the 
moving mass is sufficient to overcome several pounds pressure, and the 
tail pipe may be less than 34 feet, but to provide against any possibility 
of the water being drawn into the cylinder 
of the engine the length is made greater 
than 34 feet. The spray cone D is adjust- 
able and admits of close regulation of the 
water supply without changing the annu- 
lar form of the stream. The condensing 
water may be supplied under pressure or 
under suction. For lifts not greater than 
15 feet no supply pump is necessary, the 
water being raised by the siphon action of 
the condenser. This condenser requires 
the same amount of cooling water per 
pound of steam as the standard jet con- 
denser, and is capable of maintaining a 
vacuum of from 24 to 27 inches. A 
vacuum of 28J inches has been recorded 
for a condenser of this general type. 
(Trans. A.S.M.E., 26-388.) An atmos- 
pheric relief valve G is provided in case 
the vacuum fails from any cause, which 
will permit the steam to escape to the 
atmosphere. 

The above type of condenser is adapted 
to very muddy cooling water, since no filtration is necessary beyond the 
removal of such solid matter as may clog up the annular space H. 

In the Armour Glue Works at Chicago condensers of this type are 
successfully maintaining a 90 per cent vacuum with cooling water at 
60 degrees F. 

Siphon Condensers, Discussion: Trans. A.S.M.E., Vol. 26, p. 388. Siphon Con- 
densers: Electrical World, June, 1897, p. 818; Engr. U. S., Jan., 1906. 

249. Size of Siphon Condensers. — The size of siphon is indicated 
by the diameter of the engine exhaust pipe. 

Table 80 gives the sizes of barometric condensers as manufactured 
by prominent makers. 




Fig. 299. 



Baragwanath Siphon 
Condenser. 



476 



STEAM POWER PLANT ENGINEERING 



TABLE 80. 

SIZE OF SIPHON CONDENSERS. 



Steam to be Condensed. 


Size Usually 

Furnished, 

Inches. 


Steam to be Condensed. 


Size Usually- 


Pounds per 
Hour. 


Pounds per 
Minute. 


Pounds per Hour. 


Pounds per 
Minute. 


Furnished , 
Inches. 


2,000 
3,000 
4,000 
5,000 
6,000 


33 
50 
66 
83 
100 


5 

7 
8 
9 
9 


8,000 
10,000 
15,000 
20,000 


133 
166 
250 
333 


10 
12 

14 

14 



Vacuum 26 inches; barometer 30 inches. 

The diameter of the throat may be closely approximated by the 
empirical formula 

Diam. in inches = 0.0077 VWw, (171) 

in which 

W = weight of steam to be condensed per hour, 
w = weight of water required to condense one pound of steam. 

The maximum width of the annular opening for the admission of 
water may be obtained from the empirical formula 

Ww 



Width in inches = 



39,550 d' 



(172) 



in which 

d = diameter of the nozzle or bottom of the cone in inches. 
W and w as in equation (171). 

250. Ejector Condenser. — Fig. 300 shows a section through a Schutte 
exhaust steam "induction" condenser, illustrating the principles of the 
ejector condenser in which the momentum of flowing water ejects the 
discharge without the aid of the circulating pump. Exhaust steam 
enters the ejector through the opening marked "exhaust," passes through 
a series of inclined orifices and nozzles at considerable velocity, and, 
meeting the cooling water in the inner annular chamber, is condensed. 
The cooling water is drawn in continuously through the opening marked 
"water," by virtue of the vacuum formed, and sufficient velocity is im- 
parted to the jet to discharge the combined mass of condensed steam, 
cooling water, and air against the pressure of the atmosphere. 

Adjustment for capacity is effected by raising or lowering the ram R 
by means of the wheel H. An adjustable sleeve controls the avail- 
able area of the exhaust inlet by covering more or less openings in the 



CONDENSERS 



477 



combining tube. When the cooling water is supplied under pressure 
the openings marked " steam " and are blanked. When water is taken 
under suction and water under pressure is available for starting, is 
blanked and opening marked " steam" is connected with the pressure 
supply. When water is taken under high suction and live steam is 

used for starting, inlet marked " steam" is 
connected to live steam and an overflow 
check valve is placed at 0. Fig. 301 gives 
an outline of the necessary piping for a 
condenser installation of this type. These 
condensers are made in all sizes conforming 
with exhaust pipe diameters of. 1J to 20 
inches. The same amount of cooling water 
is required as for jet condensing and vacua 
of 20 to 25 inches are readily obtained. 




STEAM 





Fig. 300. Scnutte Ejector 
Condenser. 



Fig. 301. Piping for Schutte Ejector 
Condenser. 



251. Barometric Condensers.* — Pig. 302 shows a section through 

a Weiss counter -current condenser, illustrating the principles of a 

barometric jet condenser. The cooling water enters the upper part 

of the condensing chamber A through pipe D and falls in cascades, as 

shown in the figure, to tail pipe B, from which it flows by gravity to 

the hot well. The exhaust steam enters chamber A through pipe D, 

* The author has been informed that the word "Barometric" in connection with 
jet condensers is the registered trade mark of the Alberger Condenser Company. 



478 



STEAM POWER PLANT ENGINEERING 



and, coming in contact with the cold-water spray, is condensed. The 
air is exhausted from the top of the condenser by a dry vacuum pump 
through pipe F. In flowing to the pump the air passes upwards through 
the water spray and its temperature is lowered to that of the injection 
water, thereby reducing the volume to be exhausted. Any moisture 
passing over with the air is separated at G before reaching the air 
pump, and flows out through the small barometric tube H. The cool- 
ing water is forced to the condenser 
chamber through pipe N by any positive 
displacement pump, the actual head 
pumped against being the difference 
between the total height and that of a 
column of water corresponding to the 
degree of vacuum in the condenser. The 
main barometric tube or tail pipe B 
through which the water is discharged is 
34 feet or more in length and is provided 
with a foot valve C. The counter-current 
principle permits a much higher tempera- 
ture of hot well for the same degree of 
vacuum than does the parallel current, 
r" J [^ >T[^. U P 3 a hot-well temperature of 120 degrees 
T£ . ' p § —K : = 1 =* an d a vacuum of 27 inches being readily 

maintained. A small pipe if connecting 
the main condenser with the small baro- 
metric tube H insures at all times a 
sufficient quantity of water in the small 
auxiliary hot well to seal the tube. The 
water from this auxiliary hot well flows 
over a weir, as indicated, into a counter- 
weighted bucket M, the latter having a 
hole in the bottom which allows the nor- 
mal flow to escape. But in case a sudden 
heavy overload is thrown on the engines, 
and the adjustment is for a light load, the temperature of the discharge 
will reach the boiling point and an abnormal quantity of water will flow 
down the small barometric tube. This will cause the water to flow into 
the bucket much faster than the opening in the bottom can dispose of 
it; as a result the bucket will increase in weight and will open up a 
free-air valve L which reduces the vacuum two or three inches and 
raises the boiling point without " dropping" the vacuum entirely. E is 
the atmospheric relief valve. 




Fig. 302. 



Weiss Counter-current 
Condenser. 



CONDENSERS 



479 



Fig. 303 shows a section through the condensing chamber of an 
Alberger barometric condenser. In principles of operation the con- 
denser is similar to the Weiss, but differs considerably in details. Ex- 
haust steam enters at A and divides into two streams, one flowing 
directly to the inner chamber D, the other through the annular space E. 
Cooling water enters through B and is broken up into a fine spray by 
the serrated cone F, which is hung upon a long spring, thus auto- 
matically adjusting itself to the quantity of water entering the com 
denser. After condensing 
the exhaust steam in the 
inner cylinder the partly 
heated spray of cooling 
water in falling is brought 
in contact with the ex- 
haust steam which enters 
through the annular 
space. This process per- 
mits of a high hot-well 
temperature without af- 
fecting the degree of 
vacuum. The air which 
is not entrained by the 
cooling water and carried 
down the tail pipe col- 
lects under the spray cone 
F and ascends through 
the tubular support of 
the cone into the air 
cooler. This air cooler is 
simply a small chamber in 
which the non-conden- 
sable gases are cooled by 
a small portion of the circulating water before they are withdrawn by 
the air pump. The circulating water used for the purpose is forced 
into the cooling chamber through pipe K and falls through serrated 
openings in the bottom to the condenser proper. The air enters the 
chamber through these same openings, and is withdrawn by the air 
pump. Surrounding the cooler is a separating space of large capacity 
to allow the subsidence of any entrained moisture before the air reaches 
the vacuum pump. 

Fig. 304 shows a section through a Tomlinson type B barometric com 
denser which differs from the conventional type in the addition of an 




Fig. 303. Section through Condensing Chamber, 
Alberger Barometric Condenser. 



480 



STEAM POWER PLANT ENGINEERING 



overflow or auxiliary tail pipe. The main tail pipe takes care of the 
light loads and the overflow comes into service only on full loads and 
overloads. This arrangement reduces the quantity of circulating water 
required at light loads since it is not necessary to keep a large tail pipe 
filled with water as is the case with the single pipe design. 

Fig. 305 shows a section through a Worthington counter-current con- 
densing chamber when overhead room is not restricted, and Fig. 306 



To Dry Air Pump 




Fig. 304. 



Tomlinson Type B Barometric 
Condenser. 



Tail Pipe 
Fig. 305. 



Worthington Counter-current 
Jet Condenser. 



shows a section through the condensing chamber of a Wheeler low- 
head condenser. 

As previously outlined, surface condensers may be divided into three 
general classes, (a) water-cooled, (b) air-cooled, and (c) evaporative. 

252. Water-cooled Surface Condensers. — Water-cooled surface con- 
densers are by far the most extensive in use and only occasionally are 
the conditions such as to warrant the installation of the other class. 






CONDENSERS 



481 



They are ordinarily classified as (1) single-flow, (2) double-flow, and 
(3) multi-flow. 

Fig. 308 shows a sectional elevation through a Baragwanath vertical 
condenser, illustrating the single-flow type. It consists essentially of a 
cast-iron shell provided with two heads, into which a number of one- 
inch brass tubes are expanded. Exhaust steam fills the shell and flows 
around and between the tubes, while the cooling water is caused to 
circulate through the tubes by means of a circulating pump. The 
steam is condensed by contact with the tubes and drops to the bottom 




Fig. 306. Wheeler Low-head Centrifugal Jet Condenser. 



tube sheet, from which it is exhausted by the air pump. The circulat- 
ing water flows through the tubes in one direction only, hence the name 
" single flow." To allow for the unequal expansion of shell and tubes 
the two halves of the shell are provided with slightly thinner plates 
flanged outward, the flanges being bolted together with a spacing ring 
between them. This joint gives to the shell, in the direction of its 
length, a certain amount of elasticity which is sufficient to allow for 
the greatest possible elongation of the tubes without straining the tube 
ends and causing leakage. 

Fig. 309 shows a section through a Wheeler admiralty surface con- 
denser mounted on a combined air and circulating pump, illustrating 



482 



STEAM POWER PLANT ENGINEERING 



the typical " double-flow" surface condenser. The condenser proper 
consists of a ribbed cast-iron chamber of rectangular section fitted 

with a number of small seam- 
less drawn brass tubes through 
which the cooling water is forced 
by suitable means. The exhaust 
steam enters at the top and is 
prevented from impinging di- 
rectly against the tubes by baffle 
plates, which serve also to dis- 
tribute the steam more evenly 
over the cooling surface. The 
steam in passing between the 
tubes is condensed, and falls to 
the bottom of the chamber, 
from which it is removed, to- 




Fig. 307. General Assembly, Centrifugal Jet 
Condenser System. 



gether with the entrained 
air, by a vacuum pump. 
The water chamber be- 
tween the tube sheet and 
the head is divided into 
two compartments, as 
shown in the illustration, 
the partition being so 
arranged that the water 
flows first through the 
lower set of tubes and 
then through the upper 
set in the opposite direc- 
tion. Thus the tempera- 
ture of the cooling water 
increases as it rises, and 
reaches a maximum 
where the exhaust steam 
enters. Condensation 
begins as soon as the 
vapor enters the con- 
denser, and the surfaces 
of the tubes are at once 
covered with a thin film 
of water flowing down- 
wards from tube to tube. 




DiscliaBge 



To AitfPump 



Water 
Inlet 



Fig. 308. Baragwanath Surface Condenser. 



CONDENSER; 



483 




484 



STEAM POWER PLANT ENGINEERING 



Fig. 310 gives the details of a C. H. Wheeler & Company's surface 
condenser. The condensing chamber is of the series-parallel type in 
which the water enters the top group of tubes, then passes to the middle 
section and finally through the bottom section. Connecting chambers 
are provided at the ends of the shell as illustrated. This construction 
of water chamber keeps the condenser completely filled with cooling 
water at all times. The inlet is at the bottom but the water is carried 
up through the annular chamber to the top of the tubes. 




OBAIN TO -s^*^ COLD WATrt 

»mPUMP ,„_ 



Fig. 310. Surface-Condenser, C. H. Wheeler & Co. 



253. High-vacuum Systems. — The average reciprocating engine gives 
its best commercial economy at a vacuum of approximately 26 inches 
(referred to a 30-inch barometer), and the ordinary standard jet or sur- 
face condenser has been designed to meet this requirement. At the time 
of the introduction of the steam turbine it was discovered that a very 
high vacuum would improve turbine economies to an extent hitherto 
impossible when applied to reciprocating engines. This condition natu- 
rally created an era of development among the condenser designers. It 
became evident at once that the old types that were capable of creating 
a 26-inch or 27-inch vacuum would require considerable modification to 
maintain a vacuum of 28 inches or 29 inches. The principal improve- 
ment has been in the design of the vacuum pumps. 

Surface Condensers. — In the older types of surface condensers the 
water of condensation from the upper tubes is permitted to fall on the 
rows immediately below, thereby enveloping them with a blanket of 
water. This greatly reduces the heat transmission and necessitates 
comparatively low hot-well temperatures for a given degree of vacuum. 
By inserting baffles, Or rain plates, between the banks of tubes and 
separately draining each compartment it is possible to greatly increase 
the heat transmission and insure high hot-well temperatures. Prof. 
R. L. Weighton was the first to apply this system. (See The Efficiency 
of Surface Condensers, Proc. Inst, of Naval Architects, March, 1906.) 



CONDENSERS 



485 



Fig. 319 shows a section through a Worthington dry-tube surface con- 
denser embodying Professor Weighton's principles. 



Exhaust Inlet 




," OOOOOOOOOOOOOOOOOOOCm 

. oooooooooooooooooooooo 

/OOOOOOOOOOOOOOOOOOOOOOOl 

OOOOOOOOOOOOOO OOOOOOOOOOO- 
OOOOOOOOOOO ooooooooooo ooooooooo,/ 

ooooooooooooooooooooooooooooooo * 
ooooooooo oooooooooooooooooooooo , 

ooooooooo oooooooooooooooooooooo V 
oooooooooo oooooooooooooooooooooo 

ooooooooooooooooooooooooooooooo 

ooooooooo OOOOOOOOOOOOOOOOOOOOOO / 

oooooooooooooooooooooooooooooooy 
ooooooooo ooooooooo oooooooooooo^ 
ooooooooooooooooooo OOOOOOOOOOO / 

OOOOOOOOOOOOOOOOOOOO / 



/oooooooooo OOOOOOOOO OOOOOOOOOOO J 
> O OOOOOOOOOOO OOOOOOOOOOOOOOOOOOO 
O OOOOOOOOOOO OOOOOOOOO OOOOOOOOOO 
O OOOOOOOOOO OOOOOOOOO OOOOOOOOOOO I 

0000000000000000000000000000000/ 
. OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 
\ 000 00 000 00 OOOOOOOO 00 000 0000 OOOO , 

\OOOOO OQOOOOO OOOOOOO OOOOO / 
\, OOOOO OOOOOOO OOOOOOO OOOOOO 

^OOOOOOO OOO 00000000000000000000000 

OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 

\ OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 

\o OO OOO OO OOO OOOOOOOO OOOOOOO OOOOOOO 

OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO, 
OOOOOOOOOOOOOOOOOOOOO 




1 v — OOOOOOOOOOOOOOOOOOO 

OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 
OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 

OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 
OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 

OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 
OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 

OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 
OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 

OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 
OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 

OOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOOO 

OOOOOOOOO OOOOO OOOOOOOOOOOOOOOOOOOO / 
\ OOOOCOOOOOOOOOOOOOOOOOOOOOOOOOOOO / 

• „ 000000000000000000000000000000 xi-""j; 




Dry Air 

Outlet 



Condensed Steam Outlet 
Fig. 311. Wheeler Dry-tube Surface Condenser. 



Fig. 312 shows the general arrangement of the Worthington high- 
vacuum system. The equipment comprises a surface condenser, a 
steam-driven centrifugal pump for circulating the cooling water, a 
steam-driven rotative dry-air pump and a turbine-driven centrifugal 



486 



STEAM POWER PLANT ENGINEERING 




CONDENSERS 



487 




488 



STEAM POWER PLANT ENGINEERING 



hot-well pump. The surface condenser is piped direct to the turbine 
exhaust, only a corrugated copper expansion joint and a tee interven- 
ing. A tubular water-vapor cooler, which is in reality a small surface 
condenser, is inserted in the circulating water line between the pump 
suction and condenser, and serves to arrest all the condensable vapor 
and thus reduces the volume to be handled by the air pump. All con- 
densation, including that from the air cooler, collects in the hot well, 
from which it is pumped by a motor-driven circulating pump direct to 
heater or boiler. Cooling water is handled by a centrifugal pump hav- 
ing both suction and delivery pipes water-sealed, so that the work done 
by the pump is virtually that of overcoming the fluid friction in the 
condenser and piping. All valves and stuffing boxes are water-sealed 




Fig. 314. Parsons Vacuum Augmenter. 



to prevent any possible leakage of air, and the condenser pump cylinder 
is especially designed to avoid vapor binding. This makes it possible to 
maintain a vacuum of one-half pound absolute with cooling water at 
60 degrees F. In the high-vacuum condenser installation of the Com- 
monwealth Edison Company the dry-air pump and the circulating pump 
are direct connected to a single-cylinder Corliss engine. 

Fig. 313 shows the general arrangement of the C. H. Wheeler Com- 
pany's high-vacuum condensing outfit. The condensing chamber is 
shown in section in Fig. 310 and is described in paragraph 252. The 
wet-air pump is illustrated in Fig. 402 and is described in paragraph 324. 
No dry-air pump is needed, and the makers guarantee a vacuum within 
one inch of absolute under full-load conditions of steam turbine 
operation. 

Fig. 314 shows a section through a Parsons " vacuum augmenter" 
for increasing the vacuum in a surface condenser. A pipe is led from 



CONDENSERS 



489 



the bottom of the main condenser to an auxiliary or augmenter having 
about one-twentieth of the cooling surface of the main condenser. At 
the point indicated a small steam jet is provided which acts as an ejector 
and draws out the air and vapor from the condenser and delivers it to 
the air pump. The water seal prevents the air and vapor from return- 
ing to the condenser. With this arrangement, according to tests con- 
ducted by Mr. Parsons, if there is a vacuum of 27J or 28 inches in the 




SECTION M.-M. 
THROUGH WATER PUMP. 



Fig. 315. Westinghouse-Leblanc Multi-jet High-vacuum Condenser System. 

condenser, there may be only 26 at the air pump, which, therefore, may 
be of small size, the jet compressing the air and vapor from the con- 
denser to about one-half of its original volume. The steam jet uses about 
one and one-half per cent of the steam used by the turbine at full load. 
Jet Condensers. — Fig. 315 gives the general details of a Westing- 
house-Leblanc multi-jet condenser which, under commercial conditions, 
has realized vacua within 99 per cent of the ideal. The most striking 



490 



STEAM POWER PLANT ENGINEERING 



feature of this system lies in its compactness and simplicity, a 1500- 
kilowatt equipment being less than 9 feet in height. Referring to Fig. 
315, exhaust steam enters the condenser chamber at the upper left- 
hand opening and meets the cooling water as it is forced through spray 
nozzle C. The condensed steam and injection water fall to the bottom 
of the condenser and are removed by centrifugal pump M . The non- 
condensable vapors are withdrawn by valveless rotary air pumps P, 
through suction opening 0. Referring to section N-N through the 
air pump it will be seen that this pump consists primarily of a reverse 
Pel ton turbine wheel in conjunction with an ejector. Sealing water is 
introduced through the branch indicated by dotted outline, into the 




Fig. 316. Tomlinson Type C High-vacuum Jet Condenser. 

central chamber G, from which it passes through port H. It is then 
caught up by the blades P of the Pelton wheel, which is rotated at a 
suitable speed, and ejected into the discharge cone in the form of thin 
sheets having a high velocity. These sheets of water meet the sides 
of the discharge cone and thus form a series of water pistons, each of 
which entraps a small pocket of air and forces it out against the atmos- 
pheric pressure. In passing through the air pump the sealing water 
receives practically no increase in temperature, hence the same water 
may be used over and over again. The air pump rotor and main pump 
runner are enclosed in a common casing mounted on the same shaft. 
This arrangement makes the plant very compact and requires the use 
of only one motor to drive both pumps. There is a clear passage 
through the condenser and pump, so that should the pump stop for any 



CONDENSERS 



491 



reason air rushes into the condenser through the air pump and im- 
mediately breaks the vacuum. In starting up the condenser, steam is 
turned into auxiliary nozzle L, section N-N, for a few moments, thus 
creating sufficient vacuum to start the regular flow of water through the 
air pump. The pumps require from 1J to 3 per cent of the power 
generated by the main engines. Fig. 330 shows an application of a 
Westinghouse-Leblanc condenser to a Curtis turbine, and Fig. 331 the 
application of the Leblanc pumps to a surface condenser. 



EXHAUST TO CONDENSER . 




Fig. 317. Section through Condensing Chamber of Korting Multi-jet Condenser. 
Chamber Capable of Maintaining a Vacuum of 95 Per Cent of the Ideal without the 
use of Air Pumps. 

254. Cooling Water, Surface Condensers. — The amount of cooling 
water required per pound of steam in a surface condenser is dependent 
upon the vacuum, the temperature of the condensed steam, and the 
range in temperature of the cooling water; thus: 

H - k + 32 * 



W = 



(173) 



fa — to 
where 

II = heat content of the exhaust steam above 32 degrees F., 

h = temperature of the condensed steam, 

t = temperature of the injection water, 

t 2 = temperature of the discharge water, 

W = pounds of injection water necessary to condense one pound 
of steam. 

Example: Required the quantity of cooling water necessary to con- 
dense one pound of steam under the following conditions: Initial tem- 
perature of the cooling water 60 degrees F.; final temperature 100 
degrees F.; vacuum 26 inches, referred to 30-inch barometer. Here 



H = 1115B.t.u.,* = 60, U 



100. 



^ 1115-110 + 32 

W 100 - 60 ***' 

* See footnote, paragraph 243. 



492 



STEAM POWER PLANT ENGINEERING 



That is, the ratio of cooling water to condensed steam is approxi- 
mately 26 to 1. ' In turbine practice where vacua as high as one-half 
pound absolute are obtained, the ratio of cooling water to condensed 
steam is nearly twice this quantity. For example, if a vacuum of 
28.92 inches is desired with the barometer at 29.92 and the range of 
the circulating water temperature is 70 to 50 degrees and the tempera- 
ture of the hot well 75 degrees, the ratio will be 

1094 -75 + 32 



W 



70-50 



= 52.3. 



In 
mind 



determining the amount of cooling water it is well to bear in 
that in the ordinary condenser of the single- or double-flow type 



110 


















A Ifsw Type with Cores and Spray and 

Dry Air Pump 
B New Type without Cores and Spray 

Ordinary Pump 


130 

120 
110 
100 
90 

80 
70 
GO 
50 

Ah 




































C Old Type, Ordinary Pump 






















\A 






























B v 










*3 


















•s^C 














































3 






























\ 


\ 








Relation between Hot-Well Temperature 
and Vacuum in Surface Condensers 






\ 


\ 

\ 
\ 


































\ 
\ 
\ 
































\ 

\ 



22 23 24 25 26 27 28 29 30 

Vacuum Referred to 30 Inch Barometer 
Fig. 318. 

the temperature of the condensed steam will be from 10 to 20 degrees 
lower than that corresponding to the degree of vacuum in the con- 
denser, and that the temperature of the condensing water at the dis- 
charge point will be from 5 to 10 degrees lower than the temperature 
due to the vacuum. 

With well-designed condensers of the multi-flow type the temper- 
ature of the hot well may be from to 5 degrees lower than the tem- 
perature due to the vacuum, and the temperature of the condensing 
water at the discharge point may be equal to that due to the vacuum. 
(Proc. Inst, of Naval Arch., March, 1906.) (See Fig. 318.) 

255. Extent of Water-cooling Surface. — Theoretically, the opera- 
tion of a surface condenser is divided into two periods, (1) the period 



CONDENSERS 493 

of condensation during which the heat of vaporization at the observed 
pressure is removed and (2) the period of cooling during which the 
temperature of the condensed steam is reduced. In order to determine 
accurately the extent of cooling surface it would be necessary to cal- 
culate the heat transmission for each of the two periods. In practice, 
however, it is assumed that condensation and cooling take place simul- 
taneously, and that the mean temperature difference is a direct function 
of the temperature corresponding to the exhaust steam in the con- 
denser and that of the condensed steam and cooling water. The error 
in these assumptions has only a slight influence on the estimation of 
the cooling surface and is entirely lost' sight of in the liberal factor 
allowed in practice. 

Let S = cooling surface in square feet, 

H = heat content of the exhaust steam at condenser pressure, 
t = initial temperature of the circulating water, 
U = final temperature of the circulating water, 
h = final temperature of the condensed steam, 
t s = temperature of the exhaust steam at condenser pressure. 
U = coefficient of heat transmission, B.t.u. per hour, per degree 
difference in temperature, per square foot of cooling surface, 
d = mean difference temperature between t a and t%, and t , 
W = weight of condensed steam per hour, 



i U — to 

l0ge 



(see equation (210), Chapter XII); 



t a — ti 

and since the heat absorbed by the cooling water is equal to the heat 
given up by the steam, 

SUd = W\H -(h- 32)\, (174) 

-h 

Ud 



s = *(g-fc + *») . (175) 



Whitham (" Steam-Engine Design," p. 283) uses the arithmetic mean 
d' = t a — instead of the mean as determined from Equation (210). 

Equation (210) is based on the assumption that the fluid on each side 
of the tube is homogeneous, which is far from being true in the case of 
the air-steam mixture in a condenser, and for this reason many designers 
prefer to use the simpler arithmetic formula. 

The coefficient of heat transfer, U, as used in above equations, refers 
to the mean or average value for the entire surface since the actual heat 
transmission varies widely for different parts of the condenser; thus 



494 



STEAM POWER PLANT ENGINEERING 



the actual value of U varies from over 1000 in the first few rows of the 
tubes (where the steam comes directly into contact with the cooling 
surface) to less than 50 in the bottom row (where the tubes are practi- 
cally submerged in water of condensation) and to 3 or less for the tubes 
surrounded only by air. 




Fig. 319. Application of Weighton Dry-tube Surface Condenser to Vertical Marine Engine. 

Professor Josse of the Royal Technical School, Charlottenburg, after 
an exhaustive investigation of the subject, found that the actual value 
of U varied with 

(1) The material, thickness, shape, and cleanliness of the tube. 

(2) The velocity of the water through the tubes. 

(3) The velocity of the steam against the tubes. 

(4) The percentage of air in the steam surrounding the tubes. 

(5) The extent of submersion of the steam side of the tubes. 



CONDENSERS 



495 



Some of the results of his investigations are shown in Figs. 320 to 
322. See also Power and Engr., Feb. 2, 1909. 

The effect of thickness, material, etc., of condenser tubes is so small 
in the ultimate result and the choice and arrangement are so largely 
determined by practical consideration that they may be neglected. 

The value of U increases approximately as the square root of the 
velocity of the water flowing within the tube, so that increase in water 
velocity effects a substantial in- 
crease in the heat transmission; 
but the resistance encountered 
by the circulating water in- 
creases as the square of the 
velocity, and the power con- 
sumed in pumping the water 
increases as the third power of 
the velocity, so that a point is 
soon reached where the gain on 
the one hand may be offset by 
the loss on the other. See " The 
Transmission of Heat in Sur- 
face Condensation" by Geo. A. 
Orrok, Trans. A.S.M.E., vol. 32, 
p. 1138, 1910, for formulas per- 
taining to the value of U. This 
article contains, also, a complete 
bibliography on transmission of 
heat through tubes. 

A study of a number of in- 
stallations gave 

Old-style surface condenser, 
V = 30 to 240 feet per 

minute, average 90. 

Modern dry-tube surface condenser, 

V = 120 to 360 feet per minute, average 240. 

From the curves in Figs. 321 and 322 it will be seen that air is an 
excellent heat-insulating material; hence, the greater the amount of 
air entrained with the steam the lower will be the coefficient of heat 
transmission. The necessity of removing the air as fast as it accumu- 
lates is at once apparent. 

In the older types of surface condensers the water of condensation 
from the upper rows of tubes is permitted to fall on the rows immedi- 




12 3 4 5 6 7 

Rate of Flow of Cooling Water- Feet per Second 

Fig. 320. 



496 



STEAM POWER PLANT ENGINEERING 



"Heat Transference Tor Air 




10 20 30 40 50 60 
Sate of Flow of Air —Feet Per Second 

Fig. 321. 



70 



ately below, the water increasing in volume as it passes the successive 
banks of tubes until it completely envelops them. The coefficient U 

varies from 1000 or more in the 
upper row to less than 50 in the 
lower, giving a mean value of ap- 
proximately 250 to 350 for the 
entire surface. In estimating 
the extent of cooling surface for 
a condenser of this type an 
average figure for plain brass 
tubes with water velocities of 50 
to 100 feet per minute is U = 250. 
For a velocity of 100 to 240 feet 
per minute U may be taken 50 per 
cent greater than these figures. 
When the tubes are clean a much 
higher value may be taken, but 
a liberal factor is usually allowed 
for possible variation in the con- 
dition of operation. 
In the modern dry-tube surface condenser, designed along the lines 
of the one described in paragraph 253, in which the water of condensa- 
tion is withdrawn as rapidly as 
it is formed and air entrainment 
is reduced to a minimum, mean 
values of U = 800 to 900 are not 
unusual. In estimating the ex- 
tent of cooling surface for con- 
densers of this type an average 
value of U is 600 with water ve- 
locities of 4 to 5 feet per second. 
Example: Standard Type of 
Surface Condenser : — Required 
the number of square feet of cool- 
ing surface per i.h.p. necessary 
to condense the steam from an 
engine operating under the following conditions : Engine uses 20 pounds 
of steam per i.h.p. hour, vacuum 26 inches with barometer at 30; 
temperature of cooling water at 60 degrees. 

Here H = 1115 and t 8 = 126 (from steam tables), 
t = 60, 
k = t a - 10 = 116. 



Heat Transference for Air 



8* 



K *4 



n 

























^•j/ 
























d 


■vs 










<5 


V 


\*p 


ss^ 








f> 












/ 


K 




*.25*^ 


perSecs 






£ 















SO 



25 20 15 10 5 

Air Pressure -Inches of Mercury 

Fig. 322. 



CONDENSERS 497 

In this type of condenser average practice gives a temperature differ- 
ence of approximately 10 degrees between the temperature of the hot 
well and that corresponding to the degree of vacuum. 

t 2 = t 8 - 15 = 101. 

Any value may be fixed upon for t 2 greater than to and less than t 8 . 
The nearer ^ is to to the greater must be the quantity of circulating 
water per unit of time for a given rate of condensation. On the other 
hand, the nearer t 2 is to t 3 the less is the mean temperature difference 
d and hence the greater must be the cooling surface for a given coeffi- 
cient of heat transmission. When water is cheap and the head pumped 
against is small t 2 should be given a lower value than when water is 
costly and the discharge head is large. Average engine practice, with 
conditions as stated, gives t 2 a value of approximately 15 degrees less 
than that corresponding to the degree of vacuum. 

The logarithmic mean is, equation (210), 

j 1Q1 ~ 6Q ig i 

d ~ 126 - 60 " 42 ' 4 - 

l0ge 126 - 101 

The arithmetic mean gives 

, 1QA 60 + 101 ... 
a = 126 ~ = 45.5. 

Substitute the value of d in equation (175) and assume U = 250, the 
figure commonly used for this type of condenser. 
„ _ 20(1115-116 + 32) _ 
* " 250 X 42.4 ~ lmV% _ 

or say two square feet per i.h.p. of engine. 

Surface condensers of this type are ordinarily rated on a basis of two 
square feet per i.h.p. 

Example: Dry-tube Multi-flow Surface Condenser : — Required the 
number of square feet of cooling surface per kilowatt necessary to 
condense the steam from a steam turbine operating under the following 
conditions: Turbine uses 15 pounds of steam per kilowatt-hour; vacuum 
28.5 inches, referred to 30-inch barometer; temperature of cooling 
water 70 degrees. 

Here H = 0.9 X 1100 = 990. 

The total heat of dry steam corresponding to an absolute pressure 
of 1.5 inches is 1100, but in the case of high-vacuum turbine practice 
the steam entering the condenser is far from being dry, the quality 
varying from 0.80 to 0.95, depending upon the quality of the steam at 
admission. An average correction factor is 0.9. 

*, = 92, t = 70, h = U - 4 = 88. 



498 



STEAM POWER PLANT ENGINEERING 



In this type of condenser the hot-well temperature varies from 
= t a toti=t,-8. t 2 = t a -5 = 87. 

In this type t% varies from t 2 = t s to t 2 = t 8 — 10. 

87 - 70 



log 



92 -70 



11.5. 



92 



Arithmetic mean gives d = 92 



-87 

70 + 87 



= 13.5. 



Substitute the value of d in equation (210) and assume U = 600, the 
figure commonly used for this type of condenser. 



S = 



15 (990 - 88 + 32) 



= 2.02, 



600 X 11.5 

or say 2 square feet per kilowatt of generator. There is no standard 
rating of surface condenser for steam-turbine work because of the wide 
variation in operating conditions. A study of a number of modern 
installations gives 

1.6 to 2.5 square feet per kilowatt for large turbo-generators 

using dry-tube surface condensers. 
2.5 to 4 square feet per kilowatt for small turbo-generators 
using standard surface condensers. 
Professor Weighton found from his experiments that a surface con- 
denser constructed on the lines of the one described in paragraph 253 in 
conjunction with dry-air pumps, was capable of condensing 20 pounds 
of steam per square foot of surface per hour and maintained a vacuum 
of 28 J inches (referred to a 30-inch barometer), and this with a cooling- 
water consumption of 24 pounds per pound of condensed steam; with 
an inlet temperature of 50 degrees F. a condensation of 35 pounds of 
steam per hour per square foot of cooling surface was effected at a 
ratio of 28 pounds of cooling water per pound of steam, the vacuum 
remaining 28 J inches. See Fig. 318. (Engineering Record, May 19, 
1906, p. 615.) TABLE 81. 

EXAMPLES OF MODERN CONDENSER PROPORTIONS. 



Name of Station. 



Commonwealth Edison Co.: 

Northwest Station 

Quarry Street 

Fisk Street 

*59th St., Interborough, N. Y 

Metropolitan St. Ry., Kansas City 



Size of Turbo- 
Generators. 



20,000 
14,000 
12,000 
15,000 
10,00 



Sq. Ft. of 

Condenser 

Surface. 



32,000 
25,000 
25,000 
25,000 
22,000 



Sq. Ft. of 
Surface 
per Kw. 



1.79 
2.08 
1.67 
2.20 



* Combined Engine and Low-pressure Turbine. 



CONDENSERS 



499 



The curves in Fig. 323 are based upon equation (175) with U = 300 
and afford a simple means for determining the extent of cooling surface 
for different conditions of operation. For any other value of U multiply 
by 300 and divide by the new value of U. 

Surface Condenser Air Pumps. — See paragraphs 319 to 328. 



. 0,019 




10" 15° 20° -25° 30 o 35°i0°4550 o 60° 70 o 80 o 90l00 o 120°140lC0l80W 

Initial Temperature difference between Steam and Water 
Fig. 323. Curves for Determining the Amount of Cooling Surface. 

256. Dry-air Surface Condensers (Forced Circulation). — Where water 
is very scarce and the feed supply is reclaimed by condensing the 
exhaust steam, water-cooled condensers may be prohibitive in cost of 
operation, even when combined with cooling tower or other water-cool- 
ing device, since the latter involves a loss of water approximately 
equivalent to the amount of steam condensed, due to evaporation. 

Under these conditions air cooling has been successfully adopted. 
In the city of Kalgoorlie, West Australia, an electric station of 2000- 
horse-power capacity is equipped with air-cooled surface condensers. 



500 



STEAM POWER PLANT ENGINEERING 



The condensers have been in use five years (1906), and have given 
excellent service with very little expense and maintenance. The con- 
denser consists of a large number of narrow chambers constructed of 
thin corrugated sheet-steel plates spaced J inch between centers. Each 
chamber has 1345 square inches of cooling surface. Fifty-one of 
these chambers are grouped into a compartment and 15 compartments 
constitute a section. Each section is equipped with three motor-driven 
fans 7 feet in diameter and running normally at 320 r.p.m. In all 
there are six sections, giving a total cooling surface of 45,000 square 
feet. The steam consumption of the main engines is 16 to 16.5 pounds 
per i.h.p. hour at rated load. At full load the fans require 130 kilo- 
watts, or approximately 10 per cent of the station output. The average 
vacuum obtained is about 18 inches throughout the year and ranges 
from inches on very hot days to 22 inches in cooler weather. The 
following figures, based on actual observation, show the effect of tem- 
perature of the external air on the vacuum when condensing 32,000 
pounds of steam per hour (the rated capacity of the condenser). 



Temperature Ex- 


Vacuum, Inches 


Temperature Ex- 


Vacuum, Inches 


ternal Air, 


(referred to 30-Inch 


ternal Air, 


(referred to 30-Inch 


Degrees F. 


Barometer). 


Degrees F. 


Barometer) . 


42.8 


22 


96.8 


9.6 


50 


21.2 


100.4 


7.6 


60.8 


20 


107.6 


3.6 


68 


18.4 


113 





78.8 


16 







Air-Cooled Surface Condensers : Engineering News, Oct., 1902, p. 271 ; ibid., Vol. 
49, p. 203. 

257. Quantity of Air for Cooling (Dry-air Condenser). — The volume 
of air, under atmospheric conditions, necessary to condense steam to 
any given temperature may be determined as follows: 
Let H = heat content of the steam at condenser pressure, 
t 8 = temperature of the vapor in the condenser, 
ti = temperature of the condensed steam, 
t = temperature of the air entering condenser, 
t = temperature of the air leaving condenser, 
V — volume of air in cubic feet necessary to condense and cool 

one pound of steam, 
B = specific weight of air under atmospheric conditions, 
C = mean specific heat of air under atmospheric conditions, 
d = mean temperature difference between the air and steam, 
S = cooling surface in square feet, 

U = coefficient of heat transmission, B.t.u. per square foot per 
degree difference in temperature per hour. 



CONDENSERS 501 

Since the heat absorbed by the air must be equal to the heat given 
up by the steam, neglecting radiation, we have 

VBC (to - t) = H - h + 32, (176) 

from which 

V BC(U-t) KUi) 

For practical purposes C may be taken as the specific heat of dry air, 
the error due to this assumption being negligible even if the air is satu- 
rated with moisture. 

Example: How many cubic feet of air are necessary to condense and 
cool one pound of steam under the following conditions: Vacuum 20 
inches; temperature of entering air, leaving air, and condensed steam, 
60, 110, and 140 degrees F. respectively? 

Here H = 1130 (from steam tables), 

t = HO, h = 140, t = 60, C = 0.24, B = 0.075. 

Substituting these values in equation (177), 

V = QQ75 x 24 (110 -60) = 1135 cubic f eet of air necessar y to con ' 
dense one pound of steam under the given conditions. 

The proper area of cooling surface depends upon the value of the 
coefficient of heat transmission, which varies with the velocity and 
humidity of the air and character of the cooling surface. Accurate 
data are not available on this point. 

A few experiments made at the Armour Institute of Technology 
gave values of U = 10 to 25 B.t.u. per hour, per square foot, per degree 
difference in temperature for air velocities of 500 to 4000 feet per 
minute for corrugated-steel sheeting J inch thick. , Hence, substituting 
in equations (175) and (177) we get, for the above example, S = 1.5 
square feet of cooling surface per pound of steam condensed per hour 
for air velocity of 4000 feet per minute, and S = 3.7 square feet for a 
velocity of 500 feet per minute. 

258. Saturated-air Surface Condensers (Natural Draft). — Fig. 324 shows 
vertical and horizontal sections of a Pennel saturated-air surface con- 
denser. The apparatus consists of an upright cylindrical shell contain- 
ing a number of vertical 4-inch steel tubes through which air is drawn 
by natural draft. A centrifugal pump circulates about one half gallon 
of water per horse power per minute from a cistern below the condenser. 
The water flowing over the upper tube sheet and then descending the 
tubes by gravity forms a film over their entire interior surface. 

The condensing action is as follows: The current of exhaust steam 
entering the side of the shell at A is caused by suitable baffle plates to 



502 



STEAM POWER PLANT ENGINEERING 



circulate among the tubes, and in condensing gives up its latent heat 
to the water film, which wholly or partially evaporates, saturating the 
ascending current of air at its own temperature. The upward current 
of hot vapor-laden air carries off the heat into the atmosphere. The 
cooling water which is not evaporated and lost to the atmosphere falls 
into the cistern below to be again taken up by the circulating pump, 
the water level in the cistern being kept constant by a float governing, 
a valve on the supply pipe. The non-condensable gases collect at C> 




Fig. 324. Pennel Saturated-air Surface Condenser. 

where they are removed by the dry-air pump, while the condensed steam 
is drawn off from the bottom tube sheet by the vacuum pump and 
discharged into the hot well. An excellent feature of this device is 
that the film of water on the cooling surface is secured without inter- 
ference with the ascending air currents and also without the use of 
sprays through small orifices likely to become clogged with rust or 
sediment. Where the recovery of the condensed steam is essential and 
a high vacuum of secondary importance, condensers of this type have 
proved to be good investments on account of the low first cost. 

Table 82 gives the results of a test of a condenser of this type, taking 
steam from a 30 X 58 X 48 engine running at 45 r.p.m. (Power, 
December, 1903, p. 672; West. Elect., May 19, 1900, p. 323.) 

TABLE 82. 

TEST OF PENNEL SATURATED-AIR SURFACE CONDENSER. 

Duration of trial 9 hours 

Average steam pressure at engine by gauge 139.8 pounds 

Average vacuum, mercury column 17.5 inches 

Average temperature in condenser 123.7 degrees F. 

Average temperature of circulating water 116.4 degrees F. 



CONDENSERS 



503 



TABLE 82 (Continued). 

Average temperature of city water 52 degrees F. 

Average temperature of outside air 62 degrees F. 

Average temperature of saturated air 106 degrees F. 

Average draft in stack of condenser 1.1 inches 

Average humidity of outside air 67 per cent 

Average amount of steam condensed per hour 7950 pounds 

Avexsage amount of circulating water used per hour 114,660 pounds 

Average amount of city water used per hour 3462 pounds 

Pounds of city water per pound of steam 2.3 

Pounds of circulating water per pound of steam 14.4 

Average horse power of engine 569.7 

Steam, pounds per i.h.p. per hour 13.95 

Horse power required to run air pumps 10.5 

Horse power required to run circulating pumps 3.0 

Condensing surface, square feet 3900 

Pounds of steam condensed per square foot surface per hour 2038 

Barometer 28.58 inches 

Vapor tension corresponding to 123.7 degrees 3.82 inches 

Per cent of main engine steam used by auxiliaries 2.38 

Fig. 325 illustrates the Pennel 
"flask" type of atmospheric con- 
denser. The exhaust steam enters 
below and follows the zigzag 
course bounded by the internal 
stay channels, condensing as it 
goes and driving before it the 
non-condensable gases to the out- 
let at the top. The condensed 
steam gravitates to the bottom 
and thence to the hot well. The 
top of the flask is trough shaped 
and causes the cooling water to Fig. 325 
flow down the sides of the flask in 
a thin stream. The portion of the cooling water not evaporated collects 
at the bottom of the flask and flows to the cooling-water reservoir. 

259. Evaporative Surface Condensers. — An evaporative surface con- 
denser consists of a number of copper, brass, wrought- or cast-iron 
tubes arranged horizontally or vertically and connected to manifolds 
or chambers at each end. The exhaust steam passes through the 
tubes and a thin film of water is allowed to flow over the external 
surfaces. The cooling effect is brought about by the evaporation of 
part of the circulating water, and the general principle of operation 
is the same as that of the saturated-air condenser described above. 
Evaporation is sometimes hastened by constructing a flue over the 




Pennel Flask Type of Saturated-air 
Surface Condenser. 



504 



STEAM POWER PLANT ENGINEERING 



tubes, thereby creating a natural draft, or by means of fans. With 
horizontal cast-iron tubes and natural draft, vacua from 23 to 27 inches 
are readily maintained with a cooling surface of approximately eight 
tenths square foot per pound of steam condensed per hour. With 
vertical brass tubes and fan draft 8 pounds of steam per hour per 
square foot of cooling surface is not an unusual figure. The amount 
of cooling water evaporated per pound of steam varies from eight 
tenths to one pound, depending upon the draft. The power necessary 
to operate the pumps and fans varies from 1 to 10 per cent of the total 
output of the plant. For an interesting discussion of evaporative con- 
densers the reader is referred to the admirable article by Oldham in the 
Proceedings of the Institute of Mechanical Engineers, 1899, and re- 
produced as a serial in Engineering (London), April 28 to June 30, 1899. 
The following test of a vertical cast-iron tube evaporative surface con- 
denser (Table 83) will give some idea of the performance of this type 
of condenser. This condenser consisted of two rows of 4-inch vertical 
cast-iron pipes connected at the top by U bends and at the bottom by 
cast-iron manifolds. A perforated iron trough distributes the water 
over the center of the bend and causes it to flow in a thin stream over 
the surface of the tubes. A wet-air pump is used for withdrawing the 
condensed steam and air. No fan is used for hastening evaporation.* 

TABLE 83. 

TEST OF A CAST-IRON, VERTICAL-TUBE, EVAPORATIVE SURFACE CONDENSER 

NATURAL DRAFT. 



Date 


Sept. 12 
Wet 
29.8 
? 
272 
99 
800 
60 
1830 
600 
23.36 
117.5 
128.4 
58 

136.5 
485 
6786 
364 

1.8 

0.75 . 


Sept. 13 
Fine 


Weather 


Barometer 


29.5 


Temperature of air 


60 


Cooling surface, external 


272 


Duration of trial, minutes 


115 


Weight of steam condensed, pounds 






800 






60 




1830 




640 


Vacuum in condenser 


24.1 


Initial temperature of circulating water 
Pinal temperature of circulating water. . 




113.9 
125 




58 




131.8 


Weight of steam condensed per hour, pounds . . . 

Weight of water circulated per hour, pounds 

Weight of " make-up " water added per hour . . . 

Weight of steam condensed per square foot of 

cooling surface per hour 


427 

? 

334 

1.54 


Weight of "make-up " water per pound c 


>f steam 


0.80 







See end of paragraph 27 for evaporative surface condenser calculations. 



CONDENSERS 



505 



Evaporative Condensers: Engr., Lond., May 5, 1899, pp. 432, 442, 447; Engineer- 
ing, May 19, 1899, p. 661, June 2, 1899, p. 721, June 30, 1899, p. 861; Trans. A.S.M.E., 
14-696; Power, Nov. 16, 1909; Prac. Engr. U. S., June, 1910, p. 346. 

260. Location and Arrangement of Condensers. — In the modern 
power house one sees two general arrangements of condensers and 
auxiliaries : 

1. The independent or subdivided system, in which each engine or 
turbine is provided with its own condenser, air and circulating pumps. 

2. The central system, in which the condensers and auxiliaries are 
grouped together. Ordinarily one condenser suffices for all engines. 



In>-GtW^^ 



Dftfcliargel 





AtmosBheTe 
Atmospheric 
Relief 
Valve 



Fig. 326. Jet Condenser Located Below Engine-room Floor. 

The Independent System. — The condenser is usually placed close 
to and below the engine so that all condensation may gravitate into it. 
Figs. 326 and 329 show an application of this system with jet con- 
densers. Here each condenser receives its supply of cooling water from 
a main injection pipe and discharges into a main overflow pipe. The 
exhaust pipe leading to the condenser is by-passed through a suitable 
atmospheric relief valve to a main free exhaust header so that the 
engine may operate non-condensing in case the vacuum breaks or the 
condenser is cut out. The chief feature of this arrangement is its 
flexibility, as each unit is complete in itself and independent of the others. 



506 



STEAM POWER PLANT ENGINEERING 



By far the greater number of central stations are equipped with inde- 
pendent condensers. 

1*1 




I ' I ' I ' I ' I ' I | " | Atmospheric 
Relief Yalv.e 
Fig. 327. Jet Condenser Located Above Engine-room Floor. 

Occasionally a jet condenser is located on the same level with the 
engine or even above it, Fig. 327, but such a location should be avoided 
if possible, as it usually necessitates a larger number of bends and 




Fig. 328. Surface Condenser Located Below Engine-room Floor. 

joints in the exhaust pipes than the basement arrangement, and in- 
creases the possibility of air leakage. If the exhaust pipe does not 
drain directly into the condenser, the lowest point in the piping should 



CONDENSERS 



507 



always be provided with a drip which should be opened when the engine 
is shut down, as condensation and leakage are apt to fill the pipe with 
water if the engine stands for any length of time. The end of the drip 
should be connected so that water 
cannot be drawn back through 
the drip pipe and into the engine 
cylinder. The length of exhaust 
pipe and particularly the number 
of bends between engine and con- 
denser should be kept as small 
as possible, otherwise the engine 
may not derive the full benefit of 
the vacuum in the condenser. A 
case is recorded where the exhaust 
piping and appurtenances in con- 
nection with a 5000-horse-power 
engine caused a drop of several 
inches in vacuum between condenser and exhaust opening of the low- 
pressure cylinder. (National Engineer, December, 1906, p. 10.) The 
wet-air pump must always be located below the condenser chamber so 
that the condensation may gravitate to it. 




Fig. 329. Surface Condenser Installed in 
Connection with Pumping. 




Screens for exclusion 
of all foreign substan 
ces such as leaves, 
sticks, straw, etc 



Fig. 330. Westinghouse-Leblanc Condenser Installation Engine. 

Fig. 328 shows the arrangement of a surface condenser with com- 
bined air and circulating pump in connection with a horizontal cross 
compound engine. The condenser and appurtenances are placed below 



508 



STEAM POWER PLANT ENGINEERING 



the engine, thereby permitting the condenser to be closely connected 
to the engine. 

Discharge from Condenser 




BampBiVcnaKge. 
to Condenser 



Water Supply from 
Cold Well 

Fig. 331. Surface Condenser with Leblanc Pumps. 

Fig. 329 shows the arrangement of a surface condenser in connection 
with a pumping engine. The condenser is placed in series with the 
pump suction. 




Fig. 332. Wheeler Rectangular Jet Condenser with Centrifugal Tail Pump and Rotative 
Dry Vacuum Pump in Connection with a 10,000-kilowatt Steam Turbine. 

Central Systems. — In the central condensing systems the con- 
denser is located at any convenient point and the exhaust from all the 
engines piped to it. Any arrangement of condenser and auxiliary 
machinery may be adopted which will favor the lowest cost of installa- 
tion and expense of operation. Except where continuity of operation 
is absolutely essential, only one circulating pump and one air pump 



CONDENSERS 



509 



jBSBSfij 




510 STEAM POWER PLANT ENGINEERING 

are installed. This reduces the number of auxiliary pumps and appli- 
ances to a minimum, with a consequent decrease in first cost and main- 
tenance. With properly designed exhaust piping the condenser may 
be located at a considerable distance from the engine without undue 
loss of vacuum. 

Central condensers have found great favor in power plants in which 
the individual units are subjected to extreme variations in load, as in 
rolling mills. At the works of the Illinois Steel Company, South Chicago, 
111., one condenser takes care of the steam from 15,000 horse power of 
engines in the rail mill, and another condenses the steam from the 
15,000 horse power of engines in the Bessemer steel mill. A notable 
installation of this system in connection with street-railway work is 
in the power house of the Northwestern Elevated Company, Chicago, 
where a single condenser takes care of the exhaust steam of five engines, 
11,000 horse power in all. Fig. 333 shows the general arrangement of 
this installation. 

For a comparison of the advantages and disadvantages of the inde- 
pendent and central systems see Engineering Magazine, October, 1900, 
p. 56, Engineering, London, June 23, 1899, p. 615, and Engineering, July 
17, 1903. 

261. Power Consumption of Condenser Auxiliaries. — In estimating 

the cost of producing vacua with the different types of auxiliaries, 

steam driven, electrically driven, or belted, the power consumption is 

most conveniently expressed in terms of the equivalent heat consumption 

of the auxiliary in question and not the indicated or developed power. 

For example, suppose a power plant has a number of 1200-i.h.p. engines 

direct connected to 800-kilowatt generators and that the engines use 

20 pounds of steam per i.h.p. hour at rated load; furthermore suppose 

the engine driving the air pump (jet condenser) to indicate 24 horse 

power. Now, it is manifestly incorrect to say that the power con- 

24 
sumption of the air pump is equivalent to tkt^ = 2 per cent of the 

main engine power unless the engine driving the air pump uses 20 
pounds of steam per i.h.p. hour. As a matter of fact the small engine 
probably uses 30 to 40 pounds or more of steam per i.h.p. hour, and the 
true power consumption is 

24X30 



1200 X 20 



= 3 per cent, or more. 



If the exhaust steam is piped to the condenser, then all of this 3 per 
cent or more should be charged against the condenser; if the steam is 
piped to a heater, then only the difference between the heat entering 



CONDENSERS 



511 



the small engine and that given up to the feed water should be charged 
against it. For example, suppose the engine in the preceding examples 
uses 30 pounds of steam per i.h.p. hour when running condensing 
and 40 pounds when operating non-condensing. Let the initial steam 
pressure be 150 pounds and feed-water temperature 120 degrees F. 
when, the air pump is running condensing. If the boiler feed is not 
taken from the hot well, the heat in the exhaust steam is lost so far as 
the economy of the plant is concerned, and the heat consumption per 
i.h.p. hour is 30 {1193.6 - (120 - 32) J = 33,168 B.t.u. This repre- 
sents the cost, in heat units, of producing the vacuum, and is equivalent 
to 3 per cent of the main engine output. 



8 






















1 II 1 1 II 1 1 II 1 1 1 1 1 1 






e 
■A 




















RELATION OF POWER CONSUMPTION 

OF AUXILIARIES TO STATION 

OUTPUT 

Citizens Light, Heat and Power Co. 






1 






\ 
























\ 












6 




a 








\ 






















} 














i 

Inrnit '& of Turbine Outnut TIH.P. 






5 


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\ 




























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4 


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a 

'•A 
P-i 














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V 

■<£- 








S 




















jer 














_. 














2 
























To 


Gi 


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"Pi 


m 


r> 


















■a 
























A 


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— 
















~ 






























































— 


n 










































Jj'ull 


.Load 

L_ 













J00 200 300 400 500 
Load.E.H.P. 

Fig. 334. 



700 



If the air pump runs non-condensing and the exhaust steam is piped 
to the heater, each pound of exhaust steam gives up approximately 
950 B.t.u. per hour to the feed water and the temperature of the latter 
is raised from 120 to 180 degrees F. The heat entering the air pump 
is 40 j 1193.6 - (120 - 32) J = 44,224 B.t.u. per i.h.p. hour. But 
40 X 950 = 38,000 B.t.u. are returned to the feed water. Hence 
44,224 — 38,000 = 6224 is the net heat consumption of the air pumps 
per i.h.p. hour. This corresponds to approximately 0.55 per cent of 
the main engine output. 

In the preceding example suppose the air pump to be motor driven 

and that it requires 20 electrical horse power per hour. This will be 

20 
the equivalent of = 26.2 i.h.p. of the main engine on the 

u.oo x u.yu 

assumption that the efficiency of the small motor is 85 per cent and that 

of the engine and generator combined 90 per cent. The power required 

by the air pump will be 26.2 ■*- 1200 = 2.2 per cent of the total output. 



512 



STEAM POWER PLANT ENGINEERING 



In practice the auxiliaries use the equivalent of from 1 to 15 per cent 
of the main engine or turbine steam, depending upon the size of the plant, 
character and number of auxiliaries, and the conditions of operation. 

Table 84 gives the power consumption of the condenser auxiliaries 
in a number of installations. Fig. 334 shows the relation between the 
power consumption of the auxiliaries and the total output of the station 
at different loads for a Parsons steam-turbine installation, and Fig. 334 
shows a similar relation for a 2000-kilowatt Curtis turbine. (J. R. 
Bibbins, Power, January, 1905.) 



li 






















| 




Mil M M 


































POWER CONSUMPTION OF J 
2000 K.W. Curtis Tui 


LT7XILTARIES 








12 


ft 

=s 

n 
























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10 


















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— 

































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1500 



Load X.W. 2000 
Fig. 335. 



2500 



262. Cost of Condensers. — The following figures give an idea of 
the relative costs of the different types of condensers and auxiliaries 
for a 1000-i.h.p. plant using 20 pounds of steam per i.h.p. hour at rated 
load, or a total of 20,000 pounds per hour. Vacuum to be maintained, 
26 inches, unless otherwise stated; temperature of cooling water, 
70 degrees F.; hot-well temperature, 105 to 120 degrees F.; distance 
between engine exhaust opening and mean level of intake well, 10 feet. 

Siphon Condensers. 

1 16" siphon condenser with 6" centrifugal pump driven by 6" by 6" 

vertical engine $800 

Jet Condensers. 

1 14" by 22" by 24" jet condenser with single horizontal direct-acting 

pump 1335 

1 16" by 24" by 18" jet condenser with single vertical direct-acting 

pump 1620 

1 14" by 24" by 18" jet condenser with single vertical flywheel vacuum 

pump 1770 

1 12" by 17" by 22" by 25" jet condenser, single horizontal direct- 
acting compound pump 2200 



CONDENSERS 



513 









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514 STEAM POWER PLANT ENGINEERING 

Barometric Condensers. 

1 barometric condenser, 10" by 16" by 12" horizontal single-cylinder 
rotative dry-air pump; 8" horizontal volute centrifugal pump 
direct connected to 23-horse-power high-speed engine $2500 

1 barometric condenser, 16" by 16" dry-air pump direct connected to 
9" by 16" steam engine; positive rotary pump, for circulating 
cooling water, belted to above engine 4300 

Surface Condensers. 

1 surface condenser, 1025 square feet cooling surface, mounted over 

7§" by 14" by 14" by 12" combined air and circulating pump . . . 2100 

1 surface condenser, 1025 square feet cooling surface, with 1\" by 12" 
by 12" horizontal air pump, direct acting, and 6" centrifugal 
pump driven by 5" by 5" engine 2300 

1 surface condenser, 1025 square feet cooling surface; 5" by 12" by 
10" Edwards single-cylinder air pump and 6" centrifugal pump 
driven by a 5" by 5" engine; maximum 28", referred to 30" 
barometer 2850 

1 surface condenser, 1025 square feet cooling surface; 6" by 8" rotative 
dry-air pump; 6" by 6" Edwards wet-air pump and 6" centrifugal 
pump driven by 5" by 5" engine; maximum vacuum 29", referred 
to 30" barometer (temp, cooling water 50 degrees F.) 3500 

Westinghouse-Leblanc Jet Condenser. 

1 jet condenser with turbine-driven pumps, 20,000 pounds steam per hour, 

26" vacuum, 70 degrees F. inlet water 2150 

1 jet condenser with turbine-driven pumps, 20,000 pounds steam per 

hour, 29" vacuum, 50 degrees F. inlet water 3275 

In general the cost of complete condensing equipments installed and 
ready for operation will approximate as follows: 

Cost per Kilowatt of Main 
Generating Unit. 

Siphon condensers without air pump $2.00 to $3.00 

Jet condensers 3.00 to 4.50 

Barometric condensers with dry-air pump 4.00 to 6.00 

Surface condensers for 26-inch vacuum , 3.50 to 5.00 

High-vacuum surface condensers 3.50 to 10.00 

Leblanc jet condensers and pumps 2.00 to 6.00 

The curve in Fig. 336 shows the relative costs of complete surface 
condensing plants for steam turbines to maintain the vacua indicated. 
It will be noted how much more expensive a high-vacuum plant is than 
one designed for moderate vacua. Thus a 27-inch plant costs 25 per 
cent more than a 26-inch plant, and a 28.5-inch plant costs twice as 
much. (J. R. Bibbins, Power, January, 1905.) 

The real cost of a condensing plant, however, is not limited to the cost 
of condensing auxiliaries and piping, but should include all other costs 






CONDENSERS 



515 



26 













1 1 1 1 1 1 1 1 J i 1 1 1 1 1 i 












































































1 


































































































"55 
O 










































































4 












































^ 
































































































Comprising 
Surface Condensers^ 
Dry Air Pumps, 
Circulating Pumps, 
Jlot Well Pumps. 
Piping, Valves', etc* 




















































































Z 
































































/ 
















































/ 


















































/ 
































1 , 
















/ 






Percent of Coat of Apparatus for 26 jVactium 









100 



120 140 160 
Fig. 336. 



180 



200 



necessitated by the use of the condensing plant, including cost of extra 
building space, foundations and the like, and the attending fixed charges. 

263. Most Economical Vacuum.* — The load factor, or the ratio of 
the actual yearly load to the 
rated yearly capacity, has a 
marked. influence on the degree 
of vacuum best suited for a 
given installation, since the fixed 
charges go on whether the plant 
is running or not, while the gain 
due to the higher vacuum is 
realized only when the engines 
are operating. The higher the 
load factor the greater is the 
amount of power produced, the 
longer does the apparatus operate at best efficiency, the lower the ratio 
of fixed charges to total operating expenses, and consequently the lower 
the cost of power per unit. 

The load factor for electric-lighting stations is invariably low and 
seldom exceeds 35 per cent, with an average not far from 18 per cent. 
In street-railway work it is higher and averages about 40 per cent. In 
manufacturing plants the load factor varies considerably, but as a rule 
is somewhat higher than in either of the above cases. Tables 85 and 
86 (Power, December, 1906, p. 769) show the most economical vacua 
for different load factors for plants of 1000 kilowatts capacity with 
conditions as stated. From the tables it would seem at first glance 
that, except where coal is expensive, all the plants with low factors, 
10 per cent and under, ought to be run non-condensing. This is true 
for " one-engine" installations, but not necessarily so where there are 
a number of engines or turbines. In the latter case higher economy 
may be effected by providing only a portion of the engines with con- 
densing equipment. The engine carrying the continuous or da} 7- load 
should operate condensing, and the non-condensing engine should carry 
the peak load. In order that any of the units may be used for the day 
work, all engines could be connected to the condenser, but only those 
carrying the day load should be operated condensing. Each installa- 
tion, of course, must be considered separately and due weight given to 
the various factors entering into the problem. For an excellent article 
on the subject see " Condensers for Steam Engines and Turbines," 
Power, December, 1906, p. 769, and the Engineer, London, April 13, 
1906, p. 381; also Proc. A.S.M.E., Oct., 1910, p. 1579. 
* See also Elec'n, Lond., Jan. 14, 1910. 



516 



STEAM POWER PLANT ENGINEERING 



264. Choice of Condensers. — The proper selection of a condenser 
for a proposed installation depends upon the conditions under which 
the plant is to be operated. When there is a plentiful and cheap sup- 
ply of good condensing water suitable for boiler feed, and extremely 
high vacua are not essential, some type of jet condenser will generally 
be found most desirable. If overhead room permits, a siphon or baro- 
metric condenser will probably be most suitable and least expensive. 

TABLE 85. 

MOST ECONOMICAL VACUUM FOR STEAM TURBINES. 
Vacuum referred to 30-Inch Barometer. 





Cost of Coal, Dollars per Ton. 


Load Factor, 
per Cent. 


$1. 


50 


$2.00 


$2.50 


$3.00 


$3.50 




A 


B 


A 


B 


A 


B 


A 


B 


A 


B 


5 

10 


N.C. 

20 

24 

26.5 

27.5 

28 


N.C. 

N.C. 

17 

20 

24 

27.6 


N.C. 

23 

26.5 

27.3 

27.8 

28.2 


N.C. 

N.C. 

20 

23 

27 

27.9 


18 

25 • 

27 

27.6 

28 

28.3 


N.C. 

N.C. 

22 

25.5 

27.6 

28 


20 

26.5 

27.5 

27.8 

28.1 

28.4 


N.C. 

20 

24 

27 

27.8 

28 


22 

27 

27.7 

27.9 

28.2 

28.5 


N.C. 

22 


15 


25.8 


20 


27.5 


30 


28 


50 


28 







A. Surface-condensing plant; cost $6 per kilowatt of main generator. Fixed charges 12 per cent. 
Cost of water not included. Rated capacity of generator, 1000 kilowatts. 

B. Surface-condensing plant, including cooling towers and extra cost of land, etc.; cost $10 
per kilowatt for 26-inch plant, increasing to $14 per kilowatt for 28.5-inch plant. Fixed charges 
12 per cent. No charge for water. Rated capacity of generator, 1000 kilowatts. 

TABLE 86. 

MOST ECONOMICAL VACUUM FOR RECIPROCATING ENGINES. 

Vacuum referred to 30-Inch Barometer, 



Load Factor, 
per Cent. 


Cost of Coal, Dollars per Ton. 


$1.50 


$2.00 


$2.50 


$3.00 


$3.50 




A 


B 


A 


B 


A 


B 


A 


B 


A 


B 


10 

15 


N.C. 

16 

22.5 

24 

25.5 


N.C. 
N.C. 
N.C. 

16 

22 


15 

20 

23 

24.5 

26.7 


N.C. 

N.C. 
N.C. 
21 
23.5 


18 

22 

23.5 

25.5 

27.2 


N.C. 

N.C. 

20 

22 

23.5 


20 

22.5 

24.5 

26.4 

27.5 


N.C. 

16 
21 
23 
26.3 


22 

24 

25 

26.8 

27.7 


N.C. 
20 


20 


22 


30 


24 


50 


27 







A. Surface-condensing plant; cost $7 per kilowatt of main generator. Fixed charge 12 per cent. 
Cost of water not included. Rated capacity of generator, 1000 kilowatts. 

B. Surface-condensing plant, including cooling towers and extra cost of land, etc.; cost $11 per 
kilowatt for 26-inch plant, increasing to $13 per kilowatt for 27.5-inch plant. Other conditions 
as in A. 



CONDENSERS 517 

Where there is a plentiful supply of good water for boiler feed but 
the water which must be used for cooling purposes is very dirty the 
siphon condenser is preferable to the barometric form. A surface con- 
denser may be used in the latter case if the condensing water is not 
so dirty as to seriously impair the efficiency by coating the tubes with 
sediment, and boiler feed water is scarce. 

The air-cooled surface condenser is employed only where water of any 
kind is scarce. 

For very high vacua in connection with steam-turbine work the sur- 
face condenser is generally adopted, although the barometric condenser 
in connection with dry-air pumps and the Leblanc type of condenser 
and pumps are finding favor with many engineers. 

In selecting the type of condenser and auxiliaries due weight must 
be given to the load factor, cost of coal, water, land, building, interest, 
depreciation and the like, as outlined in the preceding paragraph. 

265. Water-Cooling Systems. — When an ample supply of cooling 
water is unobtainable, for natural or economic reasons, the circulating 
water may be used over and over again by employing suitable cooling 
devices. The three most common in practice are 

1. The simple cooling pond or tank. 

2. The spray fountain. 

3. The cooling tower. 

266. Cooling Pond. — The water is cooled partly by radiation and con- 
duction but principally by evaporation. The air is seldom saturated 
normally, and its capacity for absorbing moisture is increased on account 
of its temperature being raised by contact with the warm water and 
by radiation. The cooling action is independent of the depth of water 
and varies directly as the surface, the amount of heat dissipated for 
each square foot depending upon the temperature of the water, the rela- 
tive humidity, and the velocity of the air currents. Results of tests are 
very discordant. 

Box in his Treatise on Heat states that the pond surface should ap- 
proximate 210 square feet per nominal horse power for an engine work- 
ing twenty-four hours a day. (Treatise on Heat, Box, p. 152.) 

If the engine works only twelve hours per day, the area may be re- 
duced to 105 square feet per horse power, because the water will cool 
during the night, but in that case the depth should be such as to give a 
capacity of 300 cubic feet per horse power. These figures are based on 
a reduction in temperature of 122 to 82 degrees F., with air at 52 de- 
grees F. and humidity 85 per cent, the steam consumption per nominal 
horse power being taken at 62.5 pounds. It appears from tests that 



518 STEAM POWER PLANT ENGINEERING 

under ordinary conditions, in the northern part of the United States, with 
engines using 15 pounds of water per horse-power hour and a vacuum of 
26 inches, a reservoir having a surface of 120 square feet per horse power 
would be ample for cooling and condensing water. (W. R. Ruggles, 
Proc. A.S.M.E., April, 1912, p. 607.) 

Box gives the following formula for the rate of evaporation in per- 
fectly calm air: 

E = (243 + 3.7O(V-«0, (178) 

in which 

E = evaporation in grains per square foot per hour; 
t = temperature of the water, degrees F. ; 

V = maximum vapor tension in inches of mercury at temperature t; 

v = actual vapor tension. 

Evaporation is greatly affected by the force of the wind and varies 
from 2 to 12 times the amount determined from equation (178). 

Example: How many pounds of water will be evaporated per square 
foot per hour from a pond with the temperature of the water and air 
80 degrees F.; air perfectly calm; barometric pressure 29.5 inches and 
relative humidity 70 per cent? 

The maximum vapor tension at temperature of 80 degrees is 1.02 
inches of mercury. The actual vapor tension will be 

1.02 X 0.70 ( = relative humidity) = 0.714. 
Substitute these values in equation (178). 

E = (243 + 3.7 X 80) (1.02 - 0.714) 
= 165 grains per square foot per hour 
= 0.023 pound per square foot per hour. 

If the temperature of the water were 130 degrees F. and that of the 
surrounding air 80 degrees F., humidity 70 per cent, the evaporation 
would be 

E = (243 + 3.7 X 80) (4.5 - 0.714) 
= 2040 grains per square foot per hour 
= 0.291 pound per square foot per hour. 

Here 4.5 = maximum vapor tension, corresponding to a temperature 
of 130 degrees. 

267. Spray Fountain. — From equation (178) we see that even 
under the most favorable circumstances an enormous pond surface is 
necessary. To facilitate evaporation with a view toward reducing the 
size of the pond, the hot circulating water is sometimes distributed 
through pipes and discharged through nozzles, falling to the surface of 






CONDENSERS 



519 



the pond in a spray. The following data pertains to the spray-fountain 
installation at the power plant of the Chattanooga Electric Company, 
Chattanooga, Tenn. (Street Railway Review, March 15, 1905.) 

Adjoining the power house a pond 150 X 300 feet was excavated to 
a depth of 4 feet, the level of the water being 8 feet below the con- 
densers.* Circulating water returned from the condensers is distributed 
through a set of pipes provided with 42 nozzles through which the water 
is discharged upwards. The rectangle defined by the center lines of 
the outermost pipes is 98 feet by 125 feet. The pipes are supported on 
brick piers spaced at intervals of about 20 feet in each direction. The 




H 50 

4:0 

-90 

<D 

?80 
| 70 

I 60 

a 

K 50 
o 

l 40 

tf 30 



100 
,-90 

fe 80 

9 00 

o 

1 50 
£.40 

a 

H 30 











i i i 

Water at Nozz 


les 


r\ 












. \ 






\ 






/ 


\ > 




Water in Poi 


>4 


A 


r 


\ , 


/ 


V 


r 


.*"— * 


\> 






\ 


J 


w 


/ 


^ 


7 










r\ 






/ 


\ 




/ 


./ Air 




1 


r 


v. 


J 


\ 


A 


/ 




1 
1 


" 






\ 


/ 






1 



10 13 10 19 22 25 28 





1 






A 




rv 




-gioo 

<u 
^90 

^80 

}• 

K 60 

<o 

<> 

3 50 

<3 40 






(\ 
















vf\ 










l 




\ 


A 




\ 


\ 






\ 














/ 


J 


\ 


J 




\ 


/ 


( 


\\ 




I 




\J 


^ 


/ 


















s/ 






\ 


\ 


v 




\ 


} 




s 


n/ 












V 














v 


V 




\ 


J 


















V 














V 
















.., 



1 4 7 10 13 16 19 22 25 
Day of the Month September, 1904 



4 7 10 13 16 19 22 25 

Day .of the Month January, 1905 



Fig. 337. Curves Showing Performance of Spray Fountain; Chattanooga Electric 

Company's Power Plant. 



discharge opening of the nozzles is 1J inches in diameter, and the in- 
terior is provided with a spiral core so that in its passage the water 
is given a rotary motion, the effect of which is to greatly increase the 
spraying action. The nozzles, except on the extreme outer lines of 
piping, are placed in pairs with the axes in a vertical plane at right 
angles to the center line of the supply pipe,, the axis of each nozzle mak- 
ing an angle of 30 degrees with a vertical plane through the center of 
the supply pipe. The effect of each pair of nozzles is to throw a mass 
of spray to the height of about 15 feet, which in falling covers an area 
of 15 X 30 feet. 

A dike extending nearly across the pond near one end provides a 
canal through which the water is conducted to the suction chamber, 



520 



STEAM POWER PLANT ENGINEERING 






the object being to draw the supply from distant parts of the pond to 
give greater time for cooling. The " make-up" water is supplied by 
wells. The operation of the cooling pond for a warm month and for a 
cold month is shown in Fig. 337. Readings were taken at three-hour 
intervals. The pond supplies the circulating water for three 2000- 
square-feet Worthington surface condensers. 

Cooling Condensing Water by Means of Spray Nozzles: Power, July 1, 1908, p. 84. 

268. Cooling Towers. — A 

cooling tower consists of a 
wooden or sheet-iron housing 
open at the top and bottom 
and so arranged that the heated 
cooling water may be elevated 
to the top and distributed in 
such a manner that it falls in 
thin sheets or sprays into a 
reservoir at the bottom, air at 
the same time being drawn in 
at the bottom by natural draft 
or forced in by a fan. The 
water gives up its heat to the 
ascending current of air by 
evaporation and conduction, 
the latter, however, being a 
relatively small factor. If the 
air supply is dependent entirely 
upon convection, the system is 
known as the natural-draft or 
flue cooling tower; if the air is 
forced into the tower by fans, 
it is called a fan cooling tower. 
The different types vary prin- 
cipally in the method of water 
distribution. Fig. 338 illus- 
trates the Barnard cooling 
tower, in which the falling 
water is broken up by vertically 
suspended galvanized iron wire 
cloth mats, causing it to trickle 
in thin sheets to the bottom. A similar result is brought about in the 
Worthington tower, Fig. 339, by pieces of terra-cotta pipe 6 inches in 




DISCHARGE 
FROM 

TOWER 



Fig. 338. Barnard- Wheeler Cooling Tower. 






CONDENSERS 



521 



Tower 



diameter and two feet long placed on ends in rows. In the standard type 
of Alberger cooling tower the water trickles down the sides of swamp- 
cypress boards arranged in honeycomb fashion. In the Alberger im- 
proved type the fan is placed at the top of the tower with its shaft in a 
vertical position. The fan is operated by a Pelton water wheel which 
receives its power from a 
turbine pump. No oil lubri- 
cation is employed, and the 
operating mechanism is con- 
trolled entirely from the 
engine room. In the Jenni- 
son cooling tower the water 
is divided into a rain of 
drops, constantly retarded 
in their fall by a series of 
perforated 4 X 4-inch gal- 
vanized-iron trays arranged 
in horizontal rows and 
staggered vertically. 

With the best forms of 
cooling towers, under aver- 
age conditions, the tempera- 
ture of the circulating water 
may readily be reduced from 
40 to 50 degrees with a loss 
not exceeding 3 or 4 per 
cent of the total quantity of 
water passing through the 
tower. The power con- 
sumed by the fan in a forced- 
draft apparatus averages 2 
per cent of that developed 
by the main engines, for 
the maximum requirements 
during summer months, and 
If per cent during the 
winter. 

The location of the tower may be on the engine-room floor, on top of 
the building, or in the yard, the latter being the most adaptable. It 
may be any reasonable distance from the engine and condenser. Fig. 
340 shows a typical installation of Worthington condenser and cooling 
towers. 




c ,,,.. Hot 
^^Water 



Cold 
Water 



Suction Tank 
Fig. 339. Worthington Cooling Tower. 



522 



STEAM POWER PLANT ENGINEERING 




Rotative Dry"' 
Vacuum. Pump 



Fig. 340. Typical Cooling Tower Installation. 



CONDENSERS 523 



369. Parallel Comparison of Fan and Natural-draft Cooling Towers. 

Fan. Natural Draft. 

Size. 

Small, the forced draft providing suffi- Large draft being necessarily small, a 

cient air velocity to effect evapora- larger area must be provided to per- 

tion. form same work. 

Height limited, because loss from back Height is an advantage because the 

pressure increases with the height. tower operates on the principles of a 

Tower usually short and of large area. chimney. 

Power Consumption. 
One per cent of station output and None, 
upwards, depending upon the type of 
auxiliaries and the conditions of oper- 
ation. 

Location. 

Inside or outside. Can operate in any Outside only, unless exceptionally good 

location where sufficient head room draft is obtainable. 

and air supply are available. Preferably in the open where advantage 

Especially adapted to inconvenient loca- may be taken of prevailing winds. 

tions, as roofs, upper decks, boiler 

floors, etc. 

Conditions of Atmosphere. 

Comparatively little affected by tern- Largely affected by temperature and 

perature, considerably by humidity, humidity and wind. Draft increased 

and none by winds. by steady winds. 

Conditions of Operation. 
More especially adapted for heavy con- Especially adapted for light summer and 
tinuous duty the year round, as in heavy winter duty, as in electric-light- 

rail-plants or mills. ing plants. 

First Cost and Cost of Operation. 

First cost greater on account of mechan- First cost small by reason of simplicity 

ical construction and necessary aux- and construction. 

iliaries. First cost largely dependent upon ma- 
Cost of operation dependent upon type terials used in interior construction. 

of auxiliary and conditions of opera- Cost of operation limited to fixed 

tion. charges. 

Cooling Towers for Steam and Gas Power Plants: Trans. A.S.M.E., Vol. 31, p. 725, 
1909. 

270. Hygrometry. — The degree of saturation, or relative humidity, 
is ordinarily determined from the difference in reading of a wet- and a 
dry-bulb thermometer, thus: If the air is saturated with aqueous vapor 
no evaporation takes place from the wet bulb and the two thermometers 
give identical readings; but if it is unsaturated, evaporation occurs. 



524 



STEAM POWER PLANT ENGINEERING 



TABLE 87. 





PROPERTIES 


OF SATURATED AIR. (BAROMETER 29.921.) 








Mixture of Air Saturated with Water Vapor. 








Weight of 

1000 Cu. Ft. 

of Dry Air, 

Pounds. 


Volume of 

One Lb. of 

Dry Air, 

Cu. Ft. 


Elastic Force 

of Vapor, Ins. 

of Mercury 

(Marks & 
Davis). 


Elastic Force 
of the Dry Air 
in the Mixture, 
Ins. of Mer- 
cury. 


Weight of 1000 Cu. Ft., Lbs. 


Tem- 
pera- 
ture, de- 
grees F. 


Weight of 
the Dry 
Air, Con- 
tent. 


Weight of 

the Vapor, 

Content. 


Total 
Weight of 
the Mix- 
ture. 


1 


2 


3 


4 


5 


6 


7 


8 





86.35 


11.58 


*0.044 


29.88 


86.23 


0.081 


86.31 


10 


84.51 


11.83 


*0.069 


29.85 


84.31 


0.125 


84.43 


20 


82.75 


12.08 


*0.107 


29.81 


82.44 


0.189 


82.63 


30 


81.06 


12.33 


*0.156 


29.76 


80.62 


0.273 


80.89 


32 


80.73 


12.39 


0.180 


29.74 


80.24 


0.304 


80.54 


35 


80.24 


12.46 


0.203 


29.72 


79.70 


0.340 


80.04 


40 


79.43 


12.59 


0.248 


29.67 


78.77 


0.410 


79.18 


45 


78.64 


12.72 


0.300 


29.62 


77.86 


0.492 


78.35 


50 


77.88 


12.84 


0.362 


29.56 


76.94 


0.587 


77.53 


55 


77.12 


12.97 


0.436 


29.48 


75.98 


0.700 


76.68 


60 


76.38 


13.09 


0.522 


29.40 


75.05 


0.828 


75.88 


62 


76.08 


13.14 


0.560 


29.36 


74.66 


0.885 


75.54 


65 


75.65 


13.22 


0.622 


29.30 


74.08 


0.977 


75.06 


70 


74.94 


13.34 


0.739 


29.18 


73.08 


1.15 


74.23 


72 


74.65 


13.40 


0.790 


29.13 


72.68 


1.22 


73.90 


75 


74.24 


13.47 


0.873 


29.05 


72.08 


1.35 


73.42 


80 


73.55 


13.60 


1.03 


28.89 


71.01 


1.57 


72.58 


85 


72.87 


13.72 


1.21 


28.71 


69.92 


1.83 


71.75 


90 


72.21 


13.85 


1.42 


28.50 


68.78 


2.13 


70.91 


95 


71.56 


13.97 


1.66 


28.26 


67.59 


2.47 


70.06 


100 


70.92 


14.10 


1.93 


27.99 


66.34 


2.85 


69.19 


105 


70.29 


14.23 


2.24 


27.69 


65.05 


3.28 


68.33 


110 


69.67 


14.35 


2.59 


27.33 


63.64 


3.77 


67.41 


115 


69.07 


14.48 


2.99 


26.93 


62.16 


4.31 


66.47 


120 


68.47 


14.61 


3.44 


26.48 


60.60 


4.92 


65.52 


125 


67.88 


14.73 


3.95 


25.97 


58.92 


5.61 


64.53 


130 


67.31 


14.86 


4.52 


25.40 


57.14 


6.37 


63.51 


135 


66.74 


14.98 


5.16 


24.76 


55.23 


7.21 


62.44 


140 


66.19 


15.11 


5.88 


24.04 


53.18 


8.14 


61.32 


145 


65.64 


15.23 


6.67 


23.25 


51.01 


9.18 


60.19 


150 


65.10 


15.36 


7.57 


22.35 


48.63 


10.32 


58.95 


155 


64.57 


15.49 


8.55 


21.37 


46.12 


11.57 


57.69 


160 


64.05 


15.61 


9.65 


20.27 


43.39 


12.96 


56.35 


165 


63.54 


15.74 


10.86 


19.06 


40.47 


14.48 


54.95 


170 


63.04 


15.86 


12.20 


17.72 


37.33 


16.14 


53.47 


175 


62.54 


15.99 


13.67 


16.25 


33.96 


17.96 


51.92 


180 


62.05 


16.12 


15.29 


14.63 


30.34 


19.94 


50.28 


185 


61.57 


16.24 


17.07 


12.85 


26.44 


22.10 


48.54 


190 


61.09 


16.38 


19.02 


10.90 


22.26 


24.44 


46.70 


195 


60.63 


16.50 


21.15 


8.77 


17.17 


27.00 


44.77 


200 


60.17 


16.62 


23.47 


6.45 


12.97 


29.76 


42.73 


205 


59.71 


16.74 


26.00 


3.92 


7.82 


32.76 


40.58 


210 


59.27 


16.86 


28.76 


1.16 


2.30 


35.97 


38.27 


212 


59.09 


16.92 


29.92 








37.32 


37.32 



* Regnault. 



CONDENSERS 



525 



TABLE 87 (Continued). 





Weight of 
Water Neces- 
sary to Satu- 
rate 100 Lbs. 


Cu. Ft. of 


B.t.u. Ab- 


B.t.u, Ab- 


Cu. Ft. of 


Cu. Ft. of 




Ratio of 


Vapor from 


sorbed by 


sorbed by 


Dry Air 


Saturated 


Tem- 


Dry Air to 


One Lb. of 


1000 Cu. Ft. 


1000 Cu. Ft. 


Warmed 


Air Warmed 


pera- 


Water 


Water at Pres- 


of Dry Air 


of Saturated 


One Degree 


One Degree 


ture De- 


Vapor. 


of Dry Air. 


sure, as in 


per degree 


Air per 


F., per 


F., per 


grees F. 




Column 4. 


F.t 


Degree F.J 


B.t.u.f 


B.t.u.J 




9 


10 


11 


12 


13 


14 


15 


16 


1064.0 


0.094 




20.51 


20.52 


48.75 


48.74 





674.0 


0.148 




20.07 


20.08 


49.80 


49.79 


10 


436.0 


0.229 




19.65 


19.67 


50.89 


50.84 


20 


295.0 


0.338 




19.25 


19.27 


51.94 


51.89 


30 


264.0 


0.379 


3294.0 


19.17 


19.19 


52.16 


52.11 


32 


234.0 


0.468 


2938.0 


19.06 


19.08 


52.47 


52.41 


35 


192.0 


0.521 


2438.0 


18.86 


18.89 


53.02 


52.94 


40 


159.0 


0.632 


2038.0 


18.68 


18.72 


53.53 


53.42 


45 


131.0 


0.763 


1702.0 


18.49 


18.54 


54.04 


53.94 


50 


108.0 


0.921 


1430.0 


18.31 


18.37 


54.61 


54.43 


, 55 


91.0 


1.10 


1208.0 


18.14 


18.20 


55.12 


54.94 


60 


85.0 


1.18 


1130.0 


18.07 


18.13 


55.33 


55.16 


62 


76.0 


1.32 


1024.0 


17.96 


18.03 


55.68 


55.53 


65 


64.0 


1.57 


871.0 


17.80 


17.88 


56.18 


55.93 


70 


59.0 


1.68 


817.0 


17.74 


17.82 


56.36 


56.12 


72 


54.0 


1.87 


743.0 


17.63 


17.74 


56.71 


56.37 


75 


45.0 


2.21 


637.0 


17.47 


17.59 


57.23 


56.85 


80 


38.0 


2.62 


546.0 


17.31 


17.45 


57.77 


57.30 


85 


32.0 


3.08 


469.0 


17.14 


17.31 


58.34 


57.77 


90 


27.0 


3.65 


405.0 


17.00 


17.20 


58.82 


58.14 


95 


23.0 


4.29 


351.0 


16.84 


17.06 


59.38 


58.62 


100 


20.0 


5.04 


305.0 


16.69 


16.96 


59.92 


58.96 


105 


17.0 


5.92 


266.0 


16.54 


16.85 


60.46 


59.35 


110 


14.0 


6.93 


232.0 


16.40 


16.74 


60.97 


59.74 


115 


12.0 


8.12 


203.0 


16.26 


16.65 


61.50 


60.06 


120 


11.0 


9.46 


178.0 


16.12 


16.57 


62.03 


60.35 


125 


9.0 


11.20 


157.0 


15.98 


16.50 


62.58 


60.61 


130 


7.7 


12.90 


139.0 


15.85 


16.41 


63.09 


60.93 


135 


6.5 


15.30 


123.0 


15.72 


16.37 


63.61 


61.08 


140 


5.5 


17.90 


109.0 


15.59 


16.33 


64.14 


61.23 


145 


4.7 


21.70 


96.9 


15.46 


16.29 


64.68 


61.38 


150 


4.0 


25.10 


86.4 


15.33 


16.27 


65.23 


61.46 


155 


3.3 


29.90 


77.2 


15.21 


16.26 


65.75 


61.50 


160 


2.8 


35.80 


69.1 


15.09 


16.27 


66.27 


61.46 


165 


2.3 


43.20 


62.0 


14.97 


16.29 


66.80 


61.38 


170 


1.9 


54.80 


55.7 


14.85 


16.33 


67.34 


61.23 


175 


1.5 


65.70 


50.1 


14.74 


16.38 


67.84 


61.05 


180 


1.2 


83.80 


45.2 


14.62 


16.44 


68.40 


60.83 


185 


0.91 


111.0 


40.9 


14.51 


16.53 


68.92 


60.49 


190 


0.66 


191.0 


37.0 


14.40 


16.64 


69.44 


60.96 


195 


0.44 


229.0 


33.6 


14.29 


16.77 


69.98 


59.63 


200 


0.24 


419.0 


30.5 


14.18 


16.92 


70.52 


59.10 


205 


0.06 




27.8 


14.07 


17.09 


71.07 


58.58 


210 







26.8 


14.03 


17.16 


71.33 


58.27 


212 



t Mean specific heat of air taken as 0.2375. X Mean specific heat of water vapor taken as 0.46. 



526 STEAM POWER PLANT ENGINEERING 

The wet-bulb thermometer is thus cooled and its readings are lower than 
those of the dry-bulb. The difference in reading is a function of the 
relative humidity, and the latter may be calculated from the following 
rational psychrometric formula deduced by Willis H. Carrier (Proc. 
A.S.M.E., Nov., 1911). 

T {P-P w )d 1100 

■ . . , * " [ Pw ~ 2800- 1.3 U P? (179) 

m which L J 

h = relative humidity, per cent; 
P w = maximum tension of aqueous vapor corresponding to the tem- 
perature of the wet-bulb thermometer, inches of mercury; 
(This may be taken directly from Steam Tables.) 
P = observed barometric pressure, inches of mercury; 
d = difference in reading of the wet and dry-bulb thermometers, 

degrees F. ; 
l w = temperature of the wet-bulb thermometer; 
Pd — maximum tension of aqueous vapor corresponding to the tem- 
perature of the dry-bulb thermometer, inches of mercury. 

Exam-pie: Determine the relative humidity when the dry bulb reads 
70 degrees F., wet bulb 60 degrees F., barometer 28.0. 
From the Steam Tables we find 

P = 0.522 P = 0.739 

w 

. , r n „ 9 (28 - 0.522)10 1 100 

whence h = [0.522 - 2800 _ L3 x 60 J 5^39 

= 57.0 per cent. 

If the relative humidity h, at barometric pressure P, is to be referred 
to barometric pressure Pi, the relative humidity under the new pressure 
will be p 7, 

h = ^r- (180) 

Example: Required the relative humidity in the preceding problem 
if the air and vapor content are compressed to 30 inches 

, 30X57 ri 

hi = — ^ — =61 per cent. 

The weight of moisture per cubic foot of mixture is 

w = hy, (181) 

in which y = weight of one cubic foot of water vapor at temperature 
corresponding to the dry-bulb thermometer. (This is the density of 
steam at the specified temperature and may be taken directly from 
Steam Tables.) 



CONDENSERS 



527 



9 
3 


§ 


o 


| 


o 


o 


1 


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CONDENSERS 529 

Table 87 gives the properties of saturated air for various tempera- 
tures in terms of volumes of the mixture and is taken from the author's 
paper "Properties of Saturated and Unsaturated Air," Jour. Franklin 
Inst., Feb., 1911. 

Figs. 341 and 342 are taken from Willis H. Carrier's paper, " Rational 
Psychrometric Formulae," Proc. A.S.M.E., Nov., 1911, p. 1309, and 
give the various properties of saturated and unsaturated air for various 
temperatures and relative humidities in terms of one pound of dry air. 

271. Water-cooling Calculations. — Air is said to be completely 
saturated when it contains all the water vapor it can hold without 
causing precipitation. If the vapor content is less than that corre- 
sponding to complete saturation the air will tend to become saturated 
by absorbing moisture from surrounding objects. The drier the air 
the greater will be its affinity for moisture. The necessary latent heat 
for vaporization is supplied directly by the water producing the vapor 
or by the surrounding objects in contact with the water. Thus, in the 
open cooling tower the water vapor is absorbed from the circulating 
water, and the heat necessary to effect this vaporization is given up by 
the water, with a resultant reduction in temperature of the water itself; 
and in the evaporative surface condenser the vapor is absorbed from 
the water spray in contact with the tubes, the heat required to effect 
this vaporization being given up by the steam within the condenser 
chambers, resulting in condensation of the steam. If the air coming 
in contact with the water is very dry and at a high temperature the 
vaporization of the water may be rapid enough to cool the remaining 
water to a temperature much lower than that of the air. In this case 
practically all of the cooling is effected by evaporation. But when the 
air is at a low temperature and high relative humidity a considerable 
amount of heat may be carried away by the air by conduction. The 
quantity of air and water necessary to produce a given cooling effect 
may be determined as follows: 

Let H = total amount of heat to be abstracted, B.t.u. per hour; 
W = weight of water to be cooled, pounds per hour; 
t e = temperature of water entering cooling device; 
U = temperature of water leaving cooling device; 
U = temperature of air entering cooling device, degrees F. 

T = to + 460; 
h = temperature of air leaving cooling device, degrees F. 

T 2 = t 2 + 460; 
p = ordinary atmospheric pressure = 29.92 inches of mercury 
p a = observed atmospheric pressure, inches of mercury; 



530 STEAM POWER PLANT ENGINEERING 

p = elastic force of vapor at temperature t , inches of mercury; 

p 2 = elastic force of vapor at temperature t 2 , inches of mercury; 
Vo = volume of air entering the cooling device, cubic feet per 

hour, atmospheric conditions; 
V 2 = volume of air discharged from the cooling device at tem- 
perature t 2 ; 

d = density of dry air, at pressure p and temperature t ; 

ho = weight of moisture in 1 cubic foot of saturated air at tem- 
perature to, pounds; 

h 2 = weight of moisture in 1 cubic foot of saturated air at tem- 
perature t 2 pounds; 

zq = relative humidity of the air entering the cooling device; 

22 = relative humidity of the air leaving the cooling device; 

C = mean specific heat of dry air at constant pressure = 0.2375; * 

S = mean specific heat of water vapor at temperature t 2 ; 

r 2 = heat of vaporization at temperature t 2 . 

The pressure p\ of the dry air in atmospheric air entering the cooling 
device is 

Pi = Pa- PoZo- (182) 

The pressure pz of the dry air leaving the cooling device is 

P3 = Pa — P 2 Z 2 . (183) 

The weight w of dry air entering the cooling device under atmospheric 
temperature and pressure, pounds per hour, is 

w = ^dV . (184) 

V 

The weight w of moisture carried into the cooling device by the air, 
pounds per hour, is 

wo = hoZ Vo. (185) 

The volume of air leaving the cooling device is 

F 2 = F„^-J- (186) 

Ps J-o 

The weight w 2 of moisture carried away by the air discharged from the 
cooling device, pounds per hour, is 

W2 = h 2 Z2V 2 . (187) 

The weight of circulating water w 3 absorbed by the air in passing through 
the cooling device, pounds per hour, is 

w z = w 2 — wo. (188) 

* The mean specific heat of air varies with the temperature, see Fig. 7. 



CONDENSERS 531 

The heat H to be abstracted from the circulating water, B.t.u. per 
hour, is 

H = W (U - t e ). (189) 

The heat is dissipated, by the cooling process, in raising the tempera- 
ture of the air and vapor entering the cooling device from t to t 2 (by 
conduction) and in evaporating the moisture absorbed by the air in pass- 
ing through the apparatus (by evaporation). 

The heat H a required to raise the temperature of dry air from to to t 2) 
B.t.u. per hour, is 

H a = Cw(t 2 - to). (190) 

The heat H 8 required to superheat the water vapor entering the cool- 
ing device from temperature t to t 2 , B.t.u. per hour, is 

H 8 = w S (t 2 - t ). . (191) 

The heat H c abstracted by conduction from the circulating water, 
B.t.u. per hour, is 

H c =H a + H 8 . (192) 

The heat H e abstracted from the circulating water by evaporation, 
B.t.u. per hour, is 

H e = w,r 2 . (193) 

Though the process of evaporation is practically continued through 
the whole range in the cooling device, we are justified in using the heat 
of vaporization at the highest temperature, because the liquid was at 
this temperature entering the cooling device and the vapor is brought 
back to the temperature when leaving it. 

The total heat H t absorbed by the air in passing through the cooling 
device, B.t.u. per hour, is H t = H c + H e . (194) 

Neglecting radiation and other minor losses, the heat H t absorbed 
by the air must be equal to the heat given up by the circulating water, 

or • H t = H. (195) 

Example: Determine the quantity of air passing through the cool- 
ing tower per hour and the circulating water lost by evaporation in a 
power plant operating under the following conditions: Engines indicate 
500 horse power and consume 20 pounds steam per i.h.p. hour; tem- 
perature of the injection water, discharge water and outside air, 90, 
122, and 72 degrees F., respectively; barometer 29.5; relative humidity 
of air entering and leaving tower 70 and 90 per cent respectively; 
vacuum at condenser 25 inches. Determine also the weight of water 
evaporated in per cent of that circulated, and of the condensed steam. 



532 



STEAM POWER PLANT ENGINEERING 



the problem, 






p a = 29.5, t = 72, 


t^o = 


0.001224, 


t Po = 0.79, * t 2 = 112, 


t h 2 = 


0.003978, 


t P2 = 2.79, t e = 122, 


zo = 


0.70, 


d = 0.0747, U = 90, 


Z2 = 


0.90, 


C = 0.2375, S = 0.45, 


r 2 = 


1028.9. 



These values are obtained from Steam Tables and from Air Tables 
(Table 87). 
Substitute these values in equations (182) to (195) thus: 

(182), Vl = 29.5 - 0.79 X 0.7 

= 28.95. 
(183), p 3 = 29.5 - 2.74 X 0.9 

= 27.03. 
28.95 



(184), 
(185), 
(186), 



w 



X 0.0747 V 



29.92 
= 0.0722 V . 
w = 0.001224 X 0.7 V 
= 0.000857 Y . 

28.95 460 + 112, 



V 



27.03 460 + 72 
= 1.152 V ; 



that is, each cubic foot of dry air entering the cooling-tower is increased 
in volume to 1.152 cubic feet as it leaves. 

(187), w 2 = 0.003978 X 0.9 X 1.152 V 

= 0.004125 Vq. 
(188), w z = 0.004125 7 - 0.000857 V 

= 0.003268 7o. 

The total heat to be abstracted from the steam is 

H = 500 X 20 (1120.1 - 122 + 32) 
= 10,300,000 B.t.u. per hour. 
(189), But W (122 - 90) = 10,300,000, 
from which W = 322,000 pounds per hour. 

(190), H a = 0.2375 X 0.0722 V (112 - 72) 

= 0.6865 V . 
(191), H 8 = 0.000857 V X 0.45 (112 - 72) 
= 0.001543 Vq. 

* By assumption, U being 10 to 20 degrees lower than t e in average practice when 
the range t e — t is greater than 30 degrees, 
t Marks and Davis. 



CONDENSERS 



533 



(192), H c = 0.6865 V Q + 0.001543]F 

= 0.688 Vo. 
(193), H e = 0.003268 V X 1028.9 
= 3.365 Vo. 
H e _ 3.365 7 _ 
# c 0.688 Fo ' 

that is, the air removes 4.89 fames more heat by evaporation than by con- 
duction under the given conditions. 



(194), H t = 0.688 Vo + 3.365 V 
= 4.053 Vo. 
Hs ^ 0.001543 Vp 
H t 4.053 Vo 



= .00038; 



that is, the heat required to superheat the moisture carried into the 
tower by the air is approximately T |^ of 1 per cent of the total; hence 
an error as great as 20 per cent in the mean specific heat of the vapor is 
negligible. 

(195), 4.053 Vo = 10,300,000, 



from which 

From (188) 
Substitute 



Vo = 2,543,000 cubic feet of air per hour necessary 
to effect the required cooling. 
= 42,300 cubic feet per minute. 

w 3 = 0.003268 Vo. 

Vo = 2,543,000 in above equation. 
w z = 0.003268 X 2,543,000 

= 8320 pounds, or the weight of circulating 
water carried away per hour. 



W 



8320 
322,000 



.0258: 



that is, 2.58 per cent of the circulating water is carried away by the air 
in effecting the necessary cooling. 



w 3 



8320 



20 X 500 10,000 



= .832: 



that is, the equivalent of 83.2 per cent of the steam used by the engines 
is evaporated in the cooling tower, or the make-up water is more than 
supplied by the condensed steam. 



534 STEAM POWER PLANT ENGINEERING 

Example: Evaporation Surface Condenser. — How many cubic feet 
of air and how many pounds of water spray must be forced through an 
evaporative surface condenser of the fan type in order to condense 
1000 pounds of steam per hour and maintain a vacuum of 25 inches, 
barometer 29? (Atmospheric air 80 degrees F.. relative humidity 70 per 
cent.) The air and vapor issue from the discharge pipe under pressure 
of 4 inches of water, temperature 120 degrees F., relative humidity 
98 per cent. 

The absolute pressure in the condenser is 29.0 — 25.0 = 4 inches of 
mercury. 

The total heat to be withdrawn in order to cool and condense 1000 
pounds of steam per hour at absolute pressure of 4 inches to 120 degrees 
F. is 

1000 [1114.8 - (120 - 32)] = 1,026,000 B.t.u. 

Neglecting radiation and leakage losses, this is the heat to be ab- 
stracted per hour by the air and water spray. 

The pressure of the dry air in the mixture entering the condenser is, 
equation (182), 

Pi = 29.0 - 0.7 X 1.029 
= 28.28. 

The pressure of dry air in the mixture leaving the condenser is, 
equation (183), 

p 3 = (29.0 + 0.294) - 0.98 X 3.438 
= 25.925 

(0.294 is the value in inches of mercury of 4 inches of water-fan 
pressure). 

Let Vo = volume of atmospheric air entering the condenser. The 
volume leaving the condenser will be, equation (186), 

28.280 460 + 120 
Fs = 25\925* 460 + 80 Fo " 1A72V °' 

The weight of vapor in the condenser discharge is, equation (187), 

w 2 = 1.172 Vo X 0.004888 X 0.98 
= 0.005615 V pounds. 

The weight of vapor in the mixture entering the condenser is, equa- 
tion (185), 

wo = 0.00157 X 0.7 Vo 
= 0.001099 Vo pounds. 



CONDENSERS 535 

The amount evaporated therefore is 

w 3 = 0.005615 V - 0.001099 7 
= 0.004516 V pounds. 

The weight of dry air entering the condenser is, equation (184), 

= 0.06958 V pounds. 

The heat absorbed by the dry air in being heated from 80 degrees to 
120 degrees F. is, equation (190), 

H = Cw (t 2 - t ) 

= 0.2375 X 0.06958 V (120 - 80) 
= 0.658 7 B.t.u. 

Heat required to superheat w pounds of vapor from 80 degrees to 
120 degrees F. is, equation (191), 

#o = 0.001099 V X 0.46 (120 - 80) 
= 0.02022 V B.t.u. 

Heat absorbed by the evaporation of w^ pounds of water is, equation 
(193), 

H e = 0.004516 V X 1046.7 
= 4.720 V B.t.u. 

(Here the latent heat is taken at the lower temperature, it being the 
original temperature of the liquid.) 

Total heat absorbed by the entering air and spray is 

H t = 0.658 V + 4.720 V + 0.020 V 
= 5.398 V . 

But this represents also the heat given up by the steam, or 

5.398 V = 1,026,000, 

from which V = 190,500 cubic feet of atmospheric air necessary to 
condense and cool 1000 pounds of steam under the given conditions. 
The water spray to be injected per hour is 

0.004516 V = 0.004516 X 190,500 = 860 pounds. 

272. Test of Cooling Towers. — The following gives the results of 
a test made on the cooling-tower plant of the A. F. Brown Company 
at Elizabethport, N. J. The tower is working in connection with a 
Wheeler surface condenser of 280 square feet of cooling surface, mounted 
over a 10, 12 X 12 combined air and circulating pump. 



536 STEAM POWER PLANT ENGINEERING 

Observations made on June 24, 1904. 

Temperature of air 81 degrees 

Hygrometer 69 degrees 

Temperature of air at top of tower 89 degrees 

Temperature of water in troughs 105 degrees 

Temperature of water in tank 83 degrees 

Revolutions of fan, 239 r.p.m., air pressure.. | inch water 

Velocity of air out of tower 822 feet per minute 

Gallons of water passing over mats 385 per minute 

Vacuum 26 inches 

Temperature of air-pump discharge 87 degrees 

Observations made June 28, 1904, 9 a.m. 

Temperature of air 76 degrees 

Hygrometer 59 degrees 

Temperature of air at top of tower 81 degrees 

Temperature of water in troughs 96 degrees 

Temperature of water in tank' 78 degrees 

Revolutions of fan, 232 r.p.m., air pressure | inch water 

Velocity of air out of tower 680 feet per minute 

Gallons of water passing over mats 406 per minute 

Vacuum 25.5 inches 

Temperature of air-pump discharge 90 degrees 

Observations made June 28, 1904, 3 p.m. 

Temperature of air 74 degrees 

Hygrometer 57 degrees 

Temperature of air at top of tower 83 degrees 

Temperature of water in troughs 99 degrees 

Temperature of water in tank 80 degrees 

Revolutions of fan, 237 r.p.m., air pressure ^ inch water 

Velocity of air out of tower 769 feet per minute 

Gallons of water passing over mats 470 per minute 

Vacuum 25.5 inches 

Temperature of air-pump discharge 92 degrees 

Observations made June 29, 1904. 

Temperature of air 78 degrees 

Hygrometer 71 degrees 

Temperature of air at top of tower 86 degrees 

Temperature of water in troughs 108 degrees 

Temperature of water in tank 82 degrees 

Revolutions of fan, 241 r.p.m., air pressure f inch 

Velocity of air out of tower 772 feet per minute 

Gallons of water passing over mats 430 per minute 

Vacuum 25.5 inches 

Temperature of air-pump discharge 93 degrees 



RESULTS OF TEST OF NATURAL-DRAFT TOWER, DETROIT. 
Complete Five-Fifths Surface Installed. 
Proc. A.3.M.E.. Mid-Nov., 1909, p. 1205. 
Engines: Two 400-i.h.p. 300-kw. Macintosh & Seymour tandem-compound 

engines, overhung generators. 
Condensers: Worthington surface (admiralty type) 1600-sq. ft. reciprocating wet- 
air pump and circulating pump. 
Tower: Wood-mat construction, 24,500 sq. ft. evaporating surface, exclusive 

of shell. 
Test: March 15 to 16, 1901, 4 p.m. to 4 p m , 24 hr. 



CONDENSERS 



537 



Weather: 

Load: 
Steam : 



Water: 



Results: 



Cooling: 

Evaporation : 
Tower: 



Barometer (abs.), min 

Temperature air, deg 

Relative humidity, per cent 

600 kw. max. to 50 kw. min 

Engine efficiency = 92.5 = 875 i.h.p. max 

Weight of condensed steam per hr., lb. . . . 



A.M. P.M. AVERAGE. 

30.22 30.07; 30.14 30.27 
18.5 25; 30 25 

76 82; 58 72 

Average 244.9 kw. 

Average . .354.8 i.h.p. 
5910.6 



Temperature exhaust steam, deg. F 134 . 38 

Temperature condensed steam, deg. F 108 . 78 

Weight of steam per hour, max. load, lb 13,500 

Vacuum (abs.) 25 to 19, average about 22 

Vacuum corresponding to temperature exhaust steam. . . 25 

Vacuum possible with good condenser (10 deg. difference) 28 

Circulated per hr., lb 293,536 

Temperature hot well, average, deg. F 87.50 

Temperature cold well, average, deg. F 71.27 

Vaporization loss per hr., lb 5970 

Condenser surface per kw., sq. ft 2.66 

Steam per kw. hr., lb 24 . 3 

Steam per i.h.p. hr., lb 16.66 

Circulating water per lb. of steam, lb 49 . 6 

Steam per sq. ft. condenser surface per hr., lb 3.7 

Circulating water per sq. ft. tower surface, lb 12 

Difference in temperature between exhaust steam and 

discharge, deg. F 47 

Max. 20 deg., min. 3 deg.-5 deg. Average 16.23 

Heat dissipated per hr., B.t.u 4,769,000 

Heat per sq. ft. tower surface, B.t.u 

Heat per sq. ft. per 1000 lb. water, B.t.u 

Circulating water, per cent 

Engine steam, per cent 

Surface per kw. (average load 245 kw.), sq. ft 

Surface per kw. (max. load 600 kw.), sq. ft 

Surface per 1000 lb. steam max. load, sq. ft 

Surface per 1000 lb. steam average load, sq. ft 

Surface per 1000 lb. circulating water per deg. max. cool- 
ing, sq. ft 



195 

0.665 

2.03 

101 

100 

40.8 

1820 

4140 

4.17 





Temperature, Deg. Fahr. 


Quantities. 


Time. 


Air. 


Hot 

Well.* 


Cold 

Well. 


Water 
Cool- 
ing. 


Total 

Heat 

Head.t 


Tower 

Water, Lb. 

per Hr. 


Heat Dissi- 
pated, B.t.u. 
Lb. per Hr. 


Heat per 
Sq.Ft. Cool- 
ing Surface, 
B.t.u.perHr. 


Circulating 

Water per 

Sq. Ft., Lb. 

per Hr. 


Load, 
Kw. 


1 


2 


3 


4 


5 


6 


7 


8 


9 


10 


11 


12noon 

1.30 

2.30 
3.30 

4.30 
5.00 
6.00 
7.00 
8.00 


34 

35 

35 
35 

32.5 

28.5 

26 

24 

24 


102 

106.5 

106.5 
113 

100 

103.5 

125 

121 

123 


89 

90 

87.5 
88.5 

84 
88 
94 
94 
94.5 


13 

16.5 

19 
24.5 

16 

15.5 
31 
27 

28.5 


68 
71.5 

71.5 

78 

67.5 

75 

99 

97 

99 


375,000 

+ (375,000 

1370,200 

375,000 

375,000 

399,000 

445,500 
417,000 
427,000 
427,000 


4,880,000 

6,108,000 

7,120,000 
9,000,000 

6,384,000 

6,900,000 

12,930,000 

11,532,000 

12,174,000 


332 

415 

484 
613 

434 
470 
880 

785 

827 


25 

24.8 

25 
25 

26.6 

29.7 
27.8 
27.4 
27.4 


270 

(315 

1290 

315 

350 

365 

485 
655 
570 
600 



* Assuming a more efficient condenser, say 10 deg. difference, the probable vacuum would be 
26 deg. to 27.5 deg. This condenser actually operated at 40 deg. to 50 deg. difference. 

t Total heat head = air heating + lost head. t Difference due to rapid change in load. 



CHAPTER XII. 

FEED-WATER PURIFIERS AND HEATERS. 

273. General. — All natural waters contain more or less foreign 
matter either in suspension or solution. Waters containing carbonates 
and sulphates of magnesia and lime, soluble salts of silica, iron, and 
alumina, and suspended matter tend to form scale in the boiler and re- 
duce its steam-generating capacity and economy. The loss due to this 
cause is often overestimated but is of secondary importance to the danger 
due to retarded heat transmission which overheats and weakens the 
plates and tubes. 

Table 88 gives the results of a number of tests made on locomotive 
boiler tubes with different thicknesses and characters of scale. The 
diversity of the results indicates the futility of basing the decrease in 
conductivity on the thickness of the scale. For example, test No. 1 
shows a decrease in conductivity of 9.1 per cent for a scale 0.02 inch 
thick, while No. 16 shows a decrease of only 6.75 per cent for a scale 
over 6.5 times as thick. The scale in each case was even, hard, and 
dense. Again, No. 8 with a very soft scale 0.042 inch thick gives a 
decrease in conductivity of 9.54 per cent, whereas No. 14, also very soft 
but twice as thick, gives a decrease of only 4.95 per cent. No doubt 
the heat transmission is a function of the chemical as well as the physical 
properties, but further experiments are necessary before any specific 
conclusion can be drawn. ^ 

Waters containing acids, organic matter, and magnesium chloride 
and sulphate tend to corrode the boiler, and those containing sodium 
carbonate, organic matter, and alkalies induce priming. Even distilled 
water, as obtained from a surface condenser, is a solvent of iron to a 
certain extent and causes corrosion and pitting. Table 89 gives some 
idea of the character and extent of impurities in water from various 
localities, with an analysis of the scale produced by the water and the 
trouble in the boiler arising from its use. 

It is impossible to judge the quality of feed water merely by the 
grains of solids per gallon since a large amount of soluble salt such as 
sodium chloride will not be as deleterious as a very small amount of 
calcium sulphate. 

538 



FEED-WATER PURIFIERS AND HEATERS 



539 



TABLE 88. 

INFLUENCE OF SCALE ON HEAT TRANSMISSION. 
(Locomotive Boiler Tubes.) 



No. 


Thickness of Scale, 
Inches. 


Character of Scale. 


Decrease in Con- 
ductivity due to 
Scale. Per cent. 


1 


.02 

.02 

.033 

.033 

.038 

.04 

.04 

.042 

.047 

.065 

.07 

.07 

.085 

.089 

.11 

.13 


Hard, dense 

Hard 

Soft 

Very hard 

Medium 

Soft, porous 

Hard, dense 

Very soft 

Hard 

Medium 

Soft 

Hard 

Soft, porous 

Very soft 

Hard, porous 

Hard, dense 


9.1 


2 


2.02 


3 


4.3 


4 


3.5 


5 


4.03 


6 


6.82 


7 


3.07 


8 


9.54 • 


9 

10 


2.75 
2.39 


11 


2 38 


12 


4 43 


13 


19.0 


14 


4.95 


15. 


16 73 


16 


6.75 







From tests conducted at the University of Illinois, Railroad Gazette, Jan. 27, 1899, June 14, 
1901. See also Engineering Record, Jan. 14, 1905, p. 53; Power, February, 1903, p. 70; 
Street Railway Review, July 15, 1901, p. 415. 

The following is a rough rating according to the number of grains of 
incrusting solids per United States gallon: 

Less than 

8 grains very good. 

12 to 15 grains good. 

15 to 20 grains fair. 

20 to 30 grains bad. 

Over 30 grains very bad. 

This applies to calcium carbonate, magnesium carbonate, and mag- 
nesium chloride. For water containing sulphate of calcium and mag- 
nesium, divide the first column by 4 for the same rating. 

On account of the great variety of possible impurities the proper 
treatment to be adopted can be determined only by chemical analysis 
of the feed water in each case. 

Table 90, compiled by the Hartford Steam Boiler Inspection and 
Insurance Company, shows the number of boilers inspected by that 
company during the year 1911 and the number found defective from 
various causes. 



540 



STEAM POWER PLANT ENGINEERING 



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FEED-WATER PURIFIERS AND HEATERS 



541 



TABLE 90. 

SUMMARY OF INSPECTOR'S REPORTS FOR THE YEAR 1911. 
(Hartford Steam Boiler Inspection and Insurance Company.) 



Nature of Defects. 



Whole Number. 


Dangerous. 


19,471 


1,376 


43,663 


1,468 


2,830 


229 


13,781 


611 


9,668 


801 


2,611 


524 


5,677 


687 


7,674 


402 


3,654 


521 


5,174 


478 


565 


50 


3,225 


610 


1,204 


166 


13,015 


1,789 


9,691 


2,508 


2,009 


552 


5,956 


353 


3,402 


668 


4,436 


1,288 


430 


122 


1,209 


354 


1,334 


356 


8,145 


469 


369 


369 


9 


4 


169,202 


16,746 



Cases of deposit of sediment 

Cases of incrustation and scale. 

Cases of internal grooving 

Cases of internal corrosion 

Cases of external corrosion 

Defective braces and stays 

Settings defective 

Furnaces out of shape 

Fractured plates 

Burned plates 

Laminated plates 

Cases of defective riveting 

Defective heads 

Cases of leakage around tubes . . 

Cases of defective tubes 

Tubes too light 

Leakage of joints 

Water gauges defective 

Blow-offs defective 

Cases of deficiency of water. . . 

Safety valves overloaded 

Safety valves defective 

Pressure gages defective 

Boilers without pressure gauges 
Unclassified defects 

Condemned 



653 



The neutralization or elimination of the impurities may be effected 
by one of the following methods: 

1. Chemically. 

Boiler compounds. 
Purifying plants. 

2. Mechanically. 

Filters. 
Blow-off. 
Tube cleaners. 

3. Thermally. 

Feed-water heater. 
Distillation. 



542 



STEAM POWER PLANT ENGINEERING 



The following chart ("Boiler Waters/' W. W. Christie) outlines some 
of the troubles arising from feed water, their cause and means for 
preventing them. 



Trouble. 



Cause. 



Remedy or Palliation. 



Incrustation. 



Corrosion. 



Priming 



Sediment, mud, clay, etc.. 
Readily soluble salts 

Bicarbonate of magnesia, 
lime, iron 

Organic matter 

Sulphate of lime < 

Organic matter \ 

Grease 

Chloride or sulphate of 

magnesium 

Sugar ) 

Acid f 

Dissolved carbonic acid and S 
oxygen . £ 

Electrolytic action 

Sewage ■< 

Alkalies 

Carbonate of soda in large ) 
quantities ) 



Filtration. 
Blowing off. 
Blowing off. 

Heating feed and precipitate. 
Caustic soda. 
Lime. 
Magnesia. 
See below. 
Sodium carbonate. 
Barium chloride. 
Precipitate with alum } 
Precipitate with ferric > and filter 
chloride ) 

CaVonatTof soda } and filter 
Carbonate of soda. 

Alkali. 

Slaked lime. 

Caustic soda. 

Heating. 

Zinc plates. 

Precipitation with alum or ferric 

chloride and filter. 
Heating feed and precipitate. 

Barium chloride. 



Feed water is considered hard when the mineral matter in solution 
curdles or precipitates soap. The constituents which cause this " hard- 
ness" are carbonates and sulphates of lime, magnesia, and iron. Hard- 
ness, due to the bicarbonates, which is reduced by boiling, is said to be 
"temporary" while that which is not removed in this way is said to be 
"permanent." Low hardness, to 200 parts of calcium carbonate per 
million, is conveniently determined by means of a standard soap solu- 
tion. The latter may be obtained from chemical dealers. In deter- 
mining the degree of hardness, 50 cubic centimeters of the water are 
introduced into a 200-cubic-centimeter bottle and alcoholic soap solu- 
tion is added from a burette until a lather is obtained which covers the 
entire surface of the liquid and is permanent for five minutes. The 
degree of hardness is calculated by the use of Clark's table, thus: 



FEED-WATER PURIFIERS AND HEATERS 



543 





CLARK'S TABLE 


OF HARDNESS. 




Standard Soap 

Solution, Cubic 

Centimeters. 


Parts of CaC0 3 
Per Million. 


Standard Soap 

Solution, Cubic 

Centimeters. 


Parts of CaC0 3 
Per Million. 


Standard Soap 

Solution, cubic 

centimeters. 


Parts of CaCO-3 
Per Million. 


0.7 
1.0 
2.0 
3.0 
4.0 
5.0 


0.0 
5.0 

19.0 
32.0 
46.0 
60.0 


6.0 
7.0 
8.0 
9.0 
10.0 
11.0 


74.0 
89.0 
103.0 
118.0 
133.0 
148.0 


12.0 
13.0 
14.0 
15.0 


164.0 

180.0 
196.0 
212.0 



For waters which are harder than 200 parts per million use a solution 
ten times the strength of the standard. This method does not indicate 
the amount of reagent to be used in neutralizing the calcium sulphate 
but gives an idea of the quantity of soda crystals required to soften 
the water as far as the CaCC>3 is concerned. 

The most satisfactory way is to submit a sample of the feed water 
to a competent chemist for analysis and add the reagent recommended. 

Complete Examination of Water for Boiler Purposes: Chem. Engr., Feb., 1910, 
p. 41; Mech. Engr., Aug. 16, 1910, p. 247, May 31, 1912, p. 692; Met. and Chem. 
Engng., Jan., 1910, p. 21; Jour. Frank. Inst., Vol. CLIX, p. 217. 

Boiler Feed Water: Cassier's Mag., Oct., 1911, p. 561; June, 1910, p. 189, Jour. 
Indus, and Engng. Chem., May, 1911, p. 326. 

Boiler Corrosion: Power, June 13, 1911, p. 910, July 11, 1911, p. 67; Boiler- 
maker, Aug., 1912, p. 253; Prac. Engr., U. S., Sept. 1, 1912, p. 881. 

Boiler Scale Prevention: Elec. Wld., Jan. 6, 1910, p. 46, July 1, 1909, p. 31; Power, 
Apr. 16, 1912, p. 560, May 17, 1910, p. 888. 

274. Chemical Purification. — Chemical treatment of boiler feed 
water has been remarkably developed during recent years and a num- 
ber of manufacturing concerns make this their sole business. The 
two most common systems of chemical treatment involve (1) boiler 
compounds and (2) purifying plants. In the former the necessary 
chemical action takes place inside the boiler and in the latter the water 
is purified before it enters the boiler. In either case the usual procedure 
is to submit for analysis a sample of the feed water and the resulting 
scale to a competent chemist who will specify the character and quantity 
of chemicals necessary to bring about the desired result. 

275. Boiler Compound. — The object of treatment with boiler com- 
pounds is to neutralize the evil effects of the impurities in the feed 
water or to change them into others which are less objectionable and 
which are easily removed. When properly compounded and intro- 
duced into the boiler such preparations are of great benefit and prac- 
tically overcome the deleterious effects, but when improperly used 



544 STEAM POWER PLANT ENGINEERING 

they may produce even greater troubles than the impurities which they 
are expected to eliminate. 
Boiler compounds may be divided into three classes: 

1. Those converting the scale-forming elements into new substances 
which will not form a hard, resisting scale and which are readily removed 
by skimming, blowing off, or by tube cleaners. For example, feed 
water containing sulphates of lime and magnesia will form a dense, 
tenacious scale. If carbonate of soda be added in correct amount, the 
sulphates are converted into insoluble carbonates which are precipitated 
and form scale varying from a more or less porous, friable crust to a 
soft "mush" or mud. The resulting sulphate of soda remains in solu- 
tion and does not form scale unless allowed to concentrate, and this is 
prevented by blowing off. An excess of soda is apt to cause foaming 
and at high temperatures is liable to attack the inside of gauge glasses. 
Bisodium and trisodium phosphate, sodium tannate, fluoride of sodium, 
sugar, etc., have all proved satisfactory, but as each case requires special 
treatment no detailed discussion is possible within the scope of this 
work and the reader is referred to the accompanying bibliography. 

2. Those enveloping the newly precipitated scale-forming crystals 
with a surface which prevents them from cementing together. The 
ingredients used to bring about this result are starches, woody fibers, 
dextrine, slippery elm, and the like. 

3. Those preventing the formation of hard scale by a solvent or 
"rotting" action, as kerosene and petroleum oils. 

Boiler Compounds. — Use of Compounds: Eng. News, July 27, 1905, p. 112; 
Am. Mach., Dec. 7, 1899, p. 115, Oct. 26, 1899, p. 1014; Power, Aug., 1903; Eng. 
and Min. Jour., Aug. 12, 1905, p. 253; Elec. Wld., Oct. 7, 1909, p. 844. 

276. Use of Kerosene and Petroleum Oils in Boiler Feed Water. — 

Kerosene oil and other refined petroleum oils are sometimes used with 
good effect in boilers to prevent scale from adhering. These oils are 
said to change the deposit of lime from a hard scale to a friable material 
which may be easily removed. They are ordinarily fed to the boiler 
with the feed water, drop by drop, through a sight feed apparatus 
similar to a cylinder oil lubricator. From extended experiments made 
on a 100-horse-power tubular boiler fed with water containing 6.5 grains 
of solid matter per gallon it was found that one quart of kerosene per 
day was sufficient to keep the boiler entirely free from scale. Prior 
to the introduction of the oil the water had a corrosive action upon 
some of the fittings attached to the boiler, but after the oil had been 
used for a few months it was found that the corrosive action had ceased. 
In another case 40 gallons of kerosene were used in 24 hours in a steamer 



FEED-WATER PURIFIERS AND HEATERS 545 

of about 3000 horse power. These boilers showed no incrustation but 
considerable corrosion. Evidently oil does not have the same effect 
or give the desired results in all cases. Kerosene used in moderate 
quantities will not cause foaming. Crude oil should never be used, as 
the heavy residue causes the formation of a tough, impervious scale 
productive of bagged sheets and collapsed flues. 

Use of Kerosene in Boilers: Engr. U. S., Sept. 15, 1905, p. 634; Eng. News, May 
24, 1890, p. 497; Power, Nov. 8, 1910, p. 1993; Trans. A.S.M.E., 9-247, 11-937; 
Locomotive, July, 1890, p. 97. 

277. Use of Zinc in Boilers. — Zinc is often introduced into boilers 
to prevent corrosion. The theory is that a feeble but continuous cur- 
rent of hydrogen is generated over the whole extent of the iron by 
electrolytic action. The bubbles of hydrogen formed isolate the 
metallic surface from scale-forming substances. If there is but a little 
of the scale-forming element it is precipitated and reduced to mud; if 
there is considerable, coherent scale is produced which takes the form of 
the iron surface but does not adhere to it, being prevented from doing 
so by the intervening bubbles of hydrogen. Zinc is ordinarily sus- 
pended in the water space of the boiler in the shape of blocks, slabs, 
or as shavings in a perforated vessel. Electrical connection between 
the metallic surfaces is essential. Rolled zinc slabs 12 X 6 X i inches 
have found much favor in marine practice. Generally speaking one 
square inch of zinc surface is sufficient for every 50 pounds of water 
in the boiler, though the quantity placed in the boiler should vary with 
the hardness. The British Admiralty recommends the renewing of the 
zinc slabs whenever the decay has penetrated to a depth of J inch 
below the surface. Zinc does not prevent corrosion or scale formation 
in all cases and may even aggravate the trouble. 

Use of Zinc in Boilers: Prac. Engr., Dec., 1911, p. 835; Power, Oct. 18, 1910, 
p. 1874; Sept. 27, 1910, p. 1734. 

278. Methods of Introducing Compounds. — Boiler compounds may 
be introduced into the boiler continuously or intermittently. Small 
quantities introduced continuously or at short intervals are more effec- 
tive than large quantities at long intervals. Continuous feeding is 
ordinarily brought about by connecting the suction side of the feed 
pump with a reservoir containing the compound in solution, arranged 
similarly to an ordinary cylinder oil lubricator. In large plants an 
independent pump is often used to force the solution into the feed line. 
Intermittent feeding is brought about by temporarily connecting the 
suction of the feed pump with the reservoir containing the compound. 
The use of boiler compounds does not necessarily prevent scale from 



546 STEAM POWER PLANT ENGINEERING 

forming in time, though it will reduce the evil to a minimum. In some 
instances where compounds are used it is found necessary to run a 
tube cleaner through the tubes at certain intervals, in others such a 
course has not been found necessary. 

279. Weight of Compound Required. — The weight of compound 
introduced depends upon the nature of the reagents used and the 
character of the feed water, and ranges from a few ounces to several 
pounds per 100 gallons of feed water. For example, water containing 
4 grains of calcium sulphate and 6 grains of magnesium sulphate per 
gallon will require 3.57 pounds of carbonate of soda per 1000 gallons 
of water for the reduction of the sulphates. The chemical reaction 
and analysis is as follows: 

CaS0 4 + Na 2 C0 3 = CaC0 3 + Na 2 S0 4 , 
MgS0 4 + Na 2 G0 3 = MgC0 3 + Na 2 S0 4 . 

If x = grains of Na 2 C0 3 necessary for the calcium, 

CaS0 4 : Na 2 C0 3 + 10H 2 O = 4 : x. 
40 + 32 + 4 X 16 : 2 X 23 + 12 + 3 X 16 + 10 (2 + 16) = 4 : x. 

x = 8.41 grains. 

If y = grains of Na 2 C0 3 necessary for the magnesium, 

MgS0 4 : Na 2 C0 3 + 10H 2 O = 6 : y. 
24 + 32 + 4 X 16 : 2 X 23 + 12 + 3 X 16 + 10 (2 + 16) = 6 : y. 

y = 14.3. 

The total weight of carbonate of soda per 1000 gallons is therefore 
1000 (14.3 + 8.41) = 22,710 grains 
= 3.24 pounds. 

This amount would effect the desired result if the chemical reaction 
is permitted to take place for some time, otherwise an excess of reagent 
is necessary. As a rule, however, one quarter of the theoretical quantity 
calculated is used in boiler feed practice. 

280. Mechanical Purification. — Waters containing sand, mud, 
organic matter, and in fact all matter which is not in solution or in 
chemical combination with the water may be purified by mechanical 
filtration. Mud and sand may be eliminated by simply permitting the 
water to stand for some time in settling tanks. Suspended matter 
which will not gravitate to the bottom may be removed by filtering the 
water through coke, cloth, excelsior, or the like. Filters should be in 
duplicate for continuity of operation. 

Vegetable and other organic impurities commonly float on the surface 
of the water when the boiler is making steam, and may be blown out 
through a "surface blow-out." (See paragraph 88.) 



FEED-WATER PURIFIERS AND HEATERS 547 

Precipitated matter may be ejected from the boiler by frequent 
blowing off before it has time to adhere and bake to a crust. This 
procedure is particularly essential when boiler compounds are used. 

For description and use of mechanically operated tube cleaner see 
paragraph 92. 

281. Thermal Purification. — (See also Live Steam Purifiers, para- 
graph 298.) The carbonates of lime and magnesia are held in solution 
in fresh water by an excess of carbon dioxide and are completely pre- 
cipitated by boiling. At ordinary temperatures carbonate of lime is 
soluble in approximately 20,000 times its volume of water, at 212 
degrees F. it is slightly soluble, and at 290 degrees it is insoluble. Sul- 
phate of lime is much more soluble in cold than in hot water, and is 
completely precipitated at 290 degrees. (Revue de Mecanique, Novem- 
ber, 1901, pp. 508, 743.) Thus it will be seen that a feed heater may be. 
relied upon to remove part or all of the lime, depending upon the tem- 
perature to which the water is raised and the time in which the pre- 
cipitation is permitted to take place. 

Influence of Temperature and Concentration on the Saline Constituents of Boiler 
Water: Jour. Soc. Chem. Ind., Oct. 31, 1900, p. 885. Solubility of Sulphate of Lime: 
Rev. de Mecanique, Jan., 1901, p. 5, Nov., 1901, p. 508. 

282. Purifying Plants. — The function of a purifying plant is the 
elimination of all impurities from the feed water before it enters the 
boiler. Purifying plants are continuous or intermittent in operation 
and are modified in a number of ways to meet different conditions. 

A typical continuous system is illustrated in Fig. 343. The hard 
water enters the softener through the inlet pipe, is discharged into the 
raw water box, whence it passes over the water wheel, and thus generates 
the power necessary to maintain the reagents in constant agitation. 
From the water wheel the hard water passes into the top of the cone, 
where it meets the reagents delivered by the lift pipe and is thoroughly 
mixed with them. The reagents are dissolved in the mixing tank, 
located at the ground level, and by means of a steam, electric, or power 
pump are then elevated into the chemical tank above. One charge is 
sufficient to last ten hours or more. The reagents are apportioned 
to the amount of incoming raw water to the dividing box. (Inasmuch 
as the "head" over this stream varies directly with any fluctuation of 
the main hard water stream, the two streams are constantly maintained 
in the same proportion to each other.) In the dividing box this small 
stream is again divided by a slide which throws one part of the water 
back into the hard water stream and another part — which determines 
the rate of flow of the chemicals — into the regulating tank. As the 
level of water in the regulating tank rises, the float rises likewise and 



548 



STEAM POWER PLANT ENGINEERING 



Hard Water 
Box 




Fig. 343. Kennicott Type K Feed-water Purifier. 



FEED-WATER PURIFIERS AND HEATERS 549 

by means of a connecting chain lowers the mouth of the lift pipe in the 
reagent tank. Through this lift pipe the reagents flow into the top 
of the cone and intimately mix with the raw water. The reaction 
between the raw water and the reagents starts as soon as they meet, 
and as 'the mixture flows from the mixing plate into the reaction cone 
or downtake, the precipitation of the scale-forming and soap-destroying 
material commences to take place. Flowing at a constantly decreasing 
rate, owing to the constantly increasing diameter of the channel, the 
water passes to the bottom of the cone, turns and flows upward still at 
a constantly decreasing rate, the precipitate falling away from it as it 
moves. Finally the water passes through a filter which removes any 
slight trace of precipitate that remains; and it then is discharged from 
the top of the softener. The precipitate, which consists of the impuri- 
ties of the raw water and the softening chemicals in chemical union, 
falls to the bottom of the main tank and is from time to time discharged 
therefrom through a sludge valve. An electric indicator is provided 
which rings a bell one-half hour before a new supply of reagents is needed 
and thus notifies the attendant of the fact. The lift pipe is a tube, 
flexible for a portion of its length, through which the chemicals leave 
the chemical tank. By means of the regulating device the mouth of 
this tube is maintained at a constant depth of immersion in the surface 
of the dissolved reagents. 

In the Scaife system for water purification feed water first enters the 
heater, where it attains a temperature of from 200 to 210 degrees F. 
As a portion of the free C0 2 is driven off by the heat the carbonates of 
lime and magnesia are precipitated and are deposited in removable pans 
inside the heater. On its way the heated water is forced by the boiler 
feed pump into a large precipitating tank, where the necessary chemicals 
are introduced by two small pumps. These pumps take the solution 
of chemicals from the solution tanks which hold a sufficient quantity 
to operate the plant from eight to twelve hours. The precipitating tank 
is so constructed as to cause intimate and thorough mixing of the 
chemicals with the water. Thus the acids are neutralized, and the 
scale-forming substances are precipitated by being changed to insoluble 
substances which sink to the bottom of the precipitating tank whence 
they are readily removed. Some of the lighter substances remaining 
in suspension are carried along with the water as it passes into the 
filters, which effectively remove all suspended matter. This system 
is continuous in operation, and purification is accomplished without 
appreciably retarding the onward flow of feed water. Fig. 344 shows 
a modification of the system. The chemicals are pumped from the 
" chemical tank" into the " solution tanks," where the feed water and 



550 



STEAM POWER PLANT ENGINEERING 



chemical solution are thoroughly mixed. The treated water is taken 
from these tanks and pumped into the " precipitating tanks" where a 
large portion of the scale-forming element is precipitated. From the 
precipitating tanks the water is forced through a series of filters to 
the boiler. 



PRECIPITATING 




-J i — i L_i.L_s> L_i LJ.L.i 

Fig. 344. General Arrangement of Scaife System of Feed-water Purification. 

Fig. 345 illustrates the We-Fu-Go system of water purification. In 
this installation the water supply first enters the settling or treating 
tanks into which the chemicals are fed. A thorough mixture is effected 
by the use of the two armed paddles located near the bottom of the 
tanks. From the treating tanks the water flows by gravity into the 
filters, which remove all remaining impure solid matter which does not 
settle to the bottom of the treating tank. The pipes conducting the 
water from the settling tanks to the filter are fitted with a flexible joint 
and float so that the outlets are near the surface at all times, rising and 
falling with the water level. From the filters the purified water gravi- 
tates into the clear water storage reservoir, from which it is pumped 
into an open heater and thence to the boiler. This system is intermit- 
tent in operation, and in order to provide sufficient time for thorough 
chemical treatment of large quantities, two or more settling tanks are 
employed. Both the We-Fu-Go and Scaife systems are modified in a 
number of ways to meet different conditions. 

Fig. 346 shows the general arrangement of the Anderson system 
for preventing corrosion in condensers and removing oil from condensed 
steam. The method consists in injecting into the exhaust steam as it 
passes from the preheater to the condenser a solution containing a 
coagulant which changes the emulsion of the cylinder oil to a flaky 
condition so that it may be separated by settling, flotation, or filtering. 
The air pump delivers the water to the settling tank F, whence it is 
taken to the open gravity filters G, G, of a superficial area proportional 
to the amount of water to be passed and containing a filter bed of four 
feet of crushed quartz. This will run about four days without any 
marked difference in efficiency, after which time the bed is stirred to a 
depth of two feet by mechanical agitators and flushed with clean water, 
by which all impurities are carried to the sewer. The solution is pre- 



FEED-WATER PURIFIERS AND HEATERS 



551 



ff Treating 




rank N\ 

@ 


I^TO-Engine 
Inlet ! ! // 


Treating 


Tank X 




IfH m H '' InFi 












III \ 


&&*< 

"5^^"^ 






§ 








Outlet [/ 

"Si 

Washout \ 
"B" 


( Ta"nk I] 








Fig. 345. General Arrangement of We-Fu-Go System of Feed-water Purification. 




Fig. 346. Anderson System for Preventing Corrosion in Condensers. 



552 STEAM POWER PLANT ENGINEERING 

pared in tank A, in which the water level is preserved by a ball float 
and into which filtered water is admitted through pipe B, while the 
substance with which the water is treated is pumped in through the 
pipe D by a small pump operated from the main engine. The flow to 
the "rose head" above the condenser is controlled by the valve E, 
and a meter in this pipe records the amount being fed. The water 
ordinarily required for "make up" is sufficient to carry in the solution. 
There is very little loss of water, and the rapid corrosion of the con- 
denser tubes, which has been so great an obstacle to the successful use 
of surface condensers, is much reduced. 'The chemicals used perform 
a twofold duty, viz., to neutralize the water and make it chemically 
inactive and to coagulate the oily matter contained in the steam so 
that mechanical filtration is possible. (Power, June, 1903, p. 304.) 

Water-softening plants cost from $4 to $5 per horse power for plants 
of 1000 horse power and less, from $3 to $4 for plants of 1000 to 2000 
horse power, and as low as $1.50 for plants of 5000 horse power or more. 
The depreciation of wooden tanks is as high as 15 per cent a year, while 
that of steel tanks should not be greater than 5 per cent. Unless wooden 
tanks are considerably cheaper than steel tanks they are not a good 
investment. The cost of water purification varies from a fraction of a 
cent to 2 cents per 1000 gallons, depending upon the size of the plant 
and the quantity and character of the impurities. (American Electrician, 
March, 1905, p. 125.) 

Water Softening and Treatment for Power Plant Purposes: Chem. Engr., Jan., 
1910, p. 5; Eng. News, June 6, 1912, p. 1087; Ry. Age Gazette, Aug. 16, 1912, 
p. 288; Ry. Master Mechanic, May, 1910, p. 153; Power, May 28, 1912, p. 780; 
Apr. 18, 1911, p. 598; Prac. Engr., U. S., Mar., 1910. 

283. Economy of Preheating Feed Water. — Although a feed-water 
heater acts to some extent as a purifier its primary function is that 
of heating the feed water. Generally speaking, for every 10 degrees 
that the feed water is heated there is a gain in heat of 1 per cent and a 
corresponding saving of coal, if the heat which warms the feed water 
would otherwise be wasted. Again, the smaller the difference in tem- 
perature between the steam and the feed water the less will be the strain 
on the boiler shell due to unequal expansion and contraction, an item 
of no small consequence. 

If H represents the heat content of the steam above 32 degrees F., 
to the temperature of the cold water, and t the temperature of the water 
leaving the heater, then S, the per cent gain in heat due to heating the 
feed water, may be expressed 



FEED-WATER PURIFIERS AND HEATERS 



553 



The expression is not theoretically correct, since it assumes a con- 
stant value of unity for the specific heat, whereas the specific heat 
varies with the temperature. The variation is so slight, however, that 
it may be neglected for all practical purposes. 

Example: Steam pressure 100 pounds gauge; temperature of water 
entering heater 80 degrees F.; temperature of water leaving heater 
210 degrees F. Required, saving due to heating the feed water. 

Here H (from steam tables) is 1188, t = 80, t = 210. 

(210 - 80) 



S = 100 



32) 



1188 - (80 
= 11.4 per cent. 

This formula gives the thermal saving only, and the first cost of the 
heater, interest, depreciation, attendance, and repairs must be taken 
into consideration before the net saving measured in dollars and cents 
is ascertained. In the average installation the net saving is a sub- 
stantial one. 

Table 91 based upon formula (196) may be used in determining the 
percentages of saving due to the increase in feed-water temperature. 

TABLE 91. 

PERCENTAGE OF SAVING FOR EACH DEGREE OF INCREASE IN TEMPERATURE 

OF FEED WATER. 
(Based on Marks & Davis Steam Tables.) 



Initial 


Boiler Pressure Above Atmosphere. 


Temp, 
of Feed. 





20 


40 


60 


80 


100 


120 


140 


160 


180 


200 


32 


.0869 


.0857 


.0851 


.0846 


.0843 


.0841 


.0839 


.0837 


.0835 


.0834 


.0834 


40 


.0875 


.0863 


.0856 


.0853 


.0849 


.0846 


.0845 


.0843 


.0841 


.0840 


.0839 


50 


.0883 


.0871 


.0864 


.0859 


.0856 


.0853 


.0852 


.0850 


.0848 


.0847 


.0846 


60 


.0891 


.0878 


.0871 


.0867 


.0864 


.0861 


.0859 


.0857 


.0855 


.0854 


.0853 


70 


.0899 


.0886 


.0879 


.0874 


.0871 


.0868 


.0867 


.0865 


.0863 


.0862 


.0861 


80 


.0907 


.0894 


.0887 


.0882 


.0878 


.0876 


.0874 


.0872 


.0871 


.0870 


.0869 


90 


.0915 


.0902 


.0895 


.0890 


.0887 


.0884 


.0882 


.0880 


.0878 


.0877 


.0876 


100 


.0924 


.0910 


.0903 


.0898 


.0895 


.0892 


.0890 


.0888 


.0886 


.0885 


.0884 


110 


.0932 


.0919 


.0911 


.0906 


.0903 


.0900 


.0898 


.0896 


.0894 


.0893 


.0892 


120 


.0941 


.0927 


.0919 


.0915 


.0911 


.0908 


.0906 


.0904 


.0902 


.0901 


.0900 


130 


.0950 


.0936 


.0928 


.0923 


.0919 


.0916 


.0915 


.0912 


.0911 


.0910 


.0909 


140 


.0959 


.0945 


.0937 


.0931 


.0928 


.0925 


.0923 


.0921 


.0919 


.0918 


.0917 


150 


.0969 


.0954 


.0946 


.0940 


.0987 


.0933 


.0931 


0930 


.0928 


.0927 


.0926 


160 


.0978 


.0963 


.0955 


.0948 


.0946 


.0942 


.0940 


.0938 


.0936 


.0935 


.0934 


170 


.0988 


.0972 


.0964 


.0958 


.0955 


.0951 


.0948 


.0947 


.0945 


.0944 


.0943 


180 


.0998 


.0982 


.0973 


.0968 


.0964 


.0960 


.0958 


.0956 


.0954 


.0953 


.0952 


190 


.1008 


.0992 


.0983 


.0977 


.0973 


.0969 


.0968 


.0965 


.0964 


.0963 


.0962 


200 


.1018 


.1002 


.0993 


.0987 


.0983 


.0978 


.0977 


.0974 


.0973 


.0972 


.0971 


210 


.1029 


.1012 


.1003 


.0997 


.0993 


.0989 


.0987 


.0984 


.0983 


.0982 


.0981 


220 




.1022 


.1013 


.1007 


.1003 


.0999 


.0997 


.0994 


.0992 


.0991 


.0990 


230 




.1032 


.1023 


.1017 


.1013 


.1009 


.1007 


.1004 


.1002 


.1001 


.1000 


240 




.1043 


.1034 


.1027 


.1023 


.1019 


.1017 


.1014 


.1012 


.1011 


.1010 


250 




.1054 


.1044 


.1008 


.1034 


.1029 


.1027 


.1024 


.1022 


.1021 


.1020 



Multiply the factor in the table corresponding to any given initial temperature of feed water and boiler 
pressure by the total rise in feed-water temperature; the product will be the percentage of saving. 



554 



STEAM POWER PLANT ENGINEERING 



Feed-water Heating. — Power, June 25, 1912; Eng. News, Sept. 9, 1909, p. 284; 
Elec. Wld., March 2, 1911, p. 551; Mech. Engr., Nov. 5, 1909, p. 588; Engr. 
U. S., Jan. 1, 1906, p. 8, Aug. 15, 1904, p. 15; St. Ry. Jour., July 22, 1905, p. 145; 
Am. Elecn., Dec, 1904, p. 570; Am. Elecn., Nov., 1904; Engr., Lond., July 28, 1905. 

284. Classification of Feed-water Heaters. — Feed-water heaters may 
be classified according to the source of heat, as 

1. Exhaust steam, in which the heat is received from the exhaust of 
engines, pumps, etc. 

2. Flue gas, in which the waste chimney gases are the source of the 
heat. 

3. Live steam purifiers, or those using steam at boiler pressures; or 
according to the method of heat transmission, as 

1. Open heaters, in which the steam and feed water mingle and the 
steam in condensing gives up its heat directly to the water. 

2. Closed heaters, in which the steam and water are in separate 
chambers and the steam gives up its heat to the water by conduction. 

Heaters may also be classified according to the pressure of the heat- 
ing steam, as 

1. Vacuum or primary, in which the pressure is less than atmospheric 
and applies particularly to heaters utilizing the exhaust of condensing 
engines. These are always of the closed type. Open heaters in which 
the pressure is less than atmospheric are not usually classed as vacuum 
heaters. 

2. Atmospheric or secondary, in which the pressure is atmospheric 
or, literally, that corresponding to the back pressure on the engines 
and pumps. 

3. Pressure, in which the pressure corresponds to that in the boiler 
and in which the heat is used primarily for purifying purposes. 



CLASSIFICATION OF A FEW TYPICAL HEATERS. 
Open Atmospheric 



Exhaust steam 



Closed.. Ata>°sP h eric 

( Vacuum or pressure 



Flue Gas 

Live Steam Open Pressure. 



Cochrane 

Hoppes 

Stillwell 

Webster 

Wainwright ) Water 

Wheeler. ..JTube 

Otis ) Steam 

Berryman . ) Tube 
Green 
American 
Sturtevant 
Hoppes 
Baragwanath 



FEED-WATER PURIFIERS AND HEATERS 



555 



Heaters may be still further classified as 

1. Induced, in which only such steam is admitted as is induced by 
its condensation. That is, the feed water condenses the steam. This 
creates a partial vacuum which draws in more steam. 

2. Through, in which all the steam is forced through the heater 
irrespective of condensation. 

285. Open Heaters. — Fig. 347 gives a sectional view of a Cochrane 
special feed heater and receiver and is a typical example of an open 




Fig. 347. 

heater. Exhaust steam enters the heater through a fluted oil separa- 
tor as indicated, and passes out at the top, while the oily drips are 
automatically drained to waste by a suitable ventilated float. The 
feed water enters through an automatic valve and is distributed over 
a series of copper trays so arranged and constructed that the water is 
forced to fall in a finely divided stream before reaching the reservoir in 
the bottom. The steam coming in contact with the water particles 
gives up latent heat and condenses. Much of the scale-forming element 



556 



STEAM POWER PLANT ENGINEERING 



is deposited on the surface of the trays, from which it is readily removed. 
The suspended matter is eliminated by a coke filter in the bottom of 
the chamber, and the floating impurities are decanted by a skimmer 
or overflow weir. The particular heater shown in the illustration is 
especially designed for use in a steam-heating plant; i.e., besides per- 
forming all the functions of an open heater, it provides for the reception 
and heating of the condensation returned to it from the heating system. 




Fig. 348. Section Through Webster Heater. 



Fig. 348 gives a sectional view of a Webster "star vacuum" heater. 
Water enters the heater through balanced valve F, which is controlled 
by float E, and is deflected over a series of perforated copper trays T, T. 
Exhaust steam enters at A, passes through oil filter S, and, mingling 
with the finely divided streams of water, gives up its latent heat and is 
condensed. Only so much steam enters the heater as is condensed by 
the feed water. The condensed steam and feed water fall to the bottom 
of the upper chamber, maintaining a practically constant level WW. 
From this upper or heater chamber the water gravitates to the settling 



FEED-WATER PURIFIERS AND HEATERS 



557 



chamber at the bottom, through down-cast pipe CB. From the settling 
chamber the water rises through perforated screen M and filtering 
material P to the outlet 0. A large portion of the scale-forming element 
is precipitated on the trays or collects in the settling chamber at the 
bottom. 

Fig. 349 shows a section through a Hoppes open heater, illustrating 
the "pan" type. Exhaust steam enters at H, passes through oil filter 
0, and completely surrounds pans T, T. The feed water enters at B, 
and the rate of flow is regulated by valve F, which is controlled by a 




Fig. 349. Hoppes Horizontal Feed-water Heater. 



suitable float in the lower part of the chamber. The water in flowing 
over the sides and bottoms of the pans comes in direct contact with 
the steam. 

286. Combined Open Heater and Chemical Purifier. — Combined 
feed-water heaters and chemical purifiers are finding increased favor 
with engineers in many districts where the feed water is particularly 
bad. A description of the Webster combination will be found in Part II 
of the general catalogue issued by the Warren Webster Company, 
Camden, N. J. A description of the Cochrane-Sorge combined heater 
and chemical purifier will be found in the heater catalogue issued by the 
Harrison Safety Boiler Works, Philadelphia, Pa. 

287. Temperatures in Open Heaters. — The temperature to which feed 
water is raised in an open heater may be determined as follows: 



558 



STEAM POWER PLANT ENGINEERING 



Let H represent the heat content of the steam entering the heater, 
t the temperature of the water entering heater, 
t the temperature of the water leaving heater, and 
S the ratio of exhaust steam to the feed water, by weight. 

Then, allowing a loss of 10 per cent due to radiation, etc., 0.9 S 
(H — t + 32) will be the B.t.u. given up by the exhaust steam to each 
pound of feed water, and (t — t ) will be the B.t.u. absorbed by each 
pound of water. 

Therefore 0.9 S (H - t + 32) = t - t , from which 

t + 0.9 S (H + 32) 



t = 



1 + 0.9 S 



(197) 



If more steam passes through the heater than can be condensed by 
the feed water, then this equation gives t a fictitious value; in other 
words, t can never be greater than the temperature of the exhaust steam. 

Substituting t = 212, the maximum obtainable temperature with 
exhaust steam at atmospheric pressure, and solving for S, we find that 
only 17 per cent of the main engine exhaust is necessary to heat the 
feed water to a maximum. t is assumed to be 60 degrees F. 

Table 92 has been determined from this equation and gives the final 
temperatures obtainable in open heaters for various conditions of 
operation. 

TABLE 92. 

FINAL FEED-WATER TEMPERATURES. OPEN HEATER. 

(Temperature of steam, 212 degrees F.) 









Initial Temperature of Feed Water 


, Degrees F. 






40 


50 


60 


70 


80 


90 


100 


110 


120 


130 


3 

►^ 


2 


60.1 


69.9 


79.7 


89.5 


94.4 


109.2 


119.0 


128.8 


138.7 


148.5 


P 


3 


69.9 


79.6 


89.3 


90.1 


108.8 


118.6 


128.3 


138.0 


147.8 


157.5 


a 


4 


79.5 


89.1 


98.8 


108.5 


118.1 


127.8 


137.4 


147.1 


156.7 


166.4 


I s 


5 


89.0 


98.5 


108.1 


117.7 


127.2 


136.8 


146.4 


155.9 


165.5 


175.1 




6 


98.3 


107.7 


117.2 


126.7 


136.2 


145.7 


155.2 


164.7 


174.2 


183.6 


cd ~ 


7 


107.4 


116.8 


126.2 


135.6 


145.0 


154.4 


163.8 


173.2 


182.5 


192.1 


£3 


8 


116.4 


125.7 


135.0 


144.4 


153.7 


163.0 


172.4 


181.8 


191.0 


200.3 


>, 


9 


125.2 


134.5 


143.7 


153.0 


162.2 


171.5 


180.7 


190.0 


199.2 


208.5 


B 


10 


133.3 


143.1 


152.3 


161.4 


170.6 


179.8 


189.0 


198.1 


207.3 


212.0 


11 


142.5 


151.6 


160.7 


169.7 


178.9 


188.2 


197.0 


206.2 


212.0* 


212.0* 




12 


150.9 


159.9 


168.9 


177.9 


187.0 


196.0 


205.0 


212.0* 


212.0* 


212.0* 



* All of the steam not condensed. 



Example: A power plant has 1200 i.h.p. of engines using 20 pounds 
of steam per i.h.p. hour. Auxiliaries use equivalent of 10 per cent of 
main engine steam. Pressure in heater pounds gauge, temperature 



FEED-WATER PURIFIERS AND HEATERS 559 

of hot-well supply 110 degrees F. Required temperature of feed water 
leaving heater. 

Here H = 1150 (from steam tables), t = HO, S = 0.10. 

Substituting these values in (105), 

0.9 X 0.10 (1150 - t + 32) = t - 110. 

t = 198 degrees F. 

288. Pan Surface Required in Open Feed-water Heaters. — Pan or 

tray surface required varies according to the quality of the water with 
regard to both scale-making material and mud, and may be approxi- 
mated by the formula 

_, „ ., Lb. of water heated per hr. X horse power ,.,_._. 
Pan surface, sq. ft. = ■ (198) 





Vertical Type. 


Horizontal 
Type. 




118 
166 
500 


110 


Slightly muddy water, c 


155 


For clean water, c 


400 







289. Size of Shell, Open Heaters. — General proportions of open heaters 
vary considerably on account of the different arrangements of pans or 
trays, filter and oil-extracting devices. A fair idea of the size of shell 
required may be obtained by the formulas 

Area of shell = ^PT , , (199) 

a X length m feet 

Length of shell = Horse power 

a X area m square leet 

a = 2.15 very muddy water, 

a = 6 for slightly muddy water, 

a = 8 for clean water. 

The horse power in this case is obtained by dividing the weight of water 
heated per hour by the steam consumption of the engine per horse 
power per hour. 

Pans containing 2.5 square feet and less are usually made round, and 
larger sizes rectangular in plan. When circumstances will permit it 
is better to have not more than six pans in any one tier, since it is 
advisable to proportion the pans so as to obtain as low a velocity over 
each as practicable. 

Distance between trays or pans is seldom less than one-tenth the 
width for rectangular and one-fourth the diameter for round pans. 



560 



STEAM POWER PLANT ENGINEERING 



Volume of storage and settling chamber in horizontal heaters varies 
from 0.25 for good quality of water to 0.4 of the volume of the shell 
for muddy water, 0.33 being about the average. In the vertical type 
the settling chamber represents respectively 0.4 and 0.6 the volume of 
the shell with clear and muddy water. Filters occupy from 10 to 15 
per cent of the volume of the shell in the horizontal type and from 15 
to 20 per cent in the vertical type, the smaller percentage corresponding 
to clear water and the larger to muddy water or water containing a con- 
siderable quantity of impurities. 

Open Heaters: Cassier's Mag., Aug., 1903, p. 33; Engr. U. S., Jan. 1, 1906, pp. 
17, 78; St. Ry. Jour., Feb. 4, 1905, p. 227; Elec. Wld., Apr. 27, 1911, p. 1051. 



SURFACE BLOW 



EXHAUST 
TO HEATER 




EXHAUST FROM 
HE ATE a 



Fig. 350. 



Goubert Single-flow Closed 
Heater. 



290. Classification of Closed 
Heaters. — Closed heaters may 
be grouped into two classes: 

1. Water tube, Fig. 350, and 

2. Steam tube, Fig. 354. 

Closed heaters, both water tube 
and steam tube may operate with 

1. Parallel currents, where the 
water and steam flow in the same 
direction, Fig. 353, or with 

2. Counter currents, where the 
water and steam flow in opposite 
directions, Fig. 352. 



TUBt SHEET 




Fig. 351. 



Details of Expansion Joint, 
Goubert Heater. 



Water-tube heaters may be still further classified as 
1. Single-flow, in which the water flows through the heaters in one 
direction only, Fig. 350. 



FEED-WATER PURIFIERS AND HEATERS 



561 



2. Multi-flow, in which the water flows back and forth a number of 
times, as in Fig. 352. 

3. Coil heater, in which the water flows through one or more coils, 
as in Fig. 353. 

291. Water-tube, Closed Heaters. — Fig. 350 shows a section through 
a feed-water heater of the single-flow straight-tube type. The tubes 






Fig. 



352. Wainwright Multi-flow 
Closed Heater. 



Fig. 353. Typical Coil Heater. 



are of plain brass and the shell of cast iron. The tubes are expanded 
into the tube sheets by a roller expander. To provide for expansion 
the upper tube sheet and water chamber are secured to the main shell 
by means of a special expansion joint the details of which are shown 
in Fig. 351. R is a ring or gasket of soft annealed copper and G, G 
two gaskets of special packing with brass wire cloth insertion. These 



562 STEAM POWER PLANT ENGINEERING 

gaskets form a flexible expansion joint between C and tube sheet D, so 
that the whole upper chamber, which is carried solely by the tubes, 
is free to move up and down as the tubes expand or contract under 
varying temperatures. 

Fig. 352 shows a section through a Wainwright heater, illustrating 
the multi-flow water-tube type. The body of the heater is of cast 
iron, the tubes of corrugated copper. The water passes through the 
tubes and the steam surrounds them. The feed water and exhaust 
steam do not mingle, and hence the oil in the exhaust does not con- 
taminate the water. The water chambers are divided into several 
compartments, as shown in the illustration, and the partitions are so 
arranged that the flow of feed water is directed back and forth through 
the various groups of tubes in succession. This arrangement gives a 
higher velocity of flow than the non-return type of heater, and therefore 
increases the rate of heat absorption. The mud and impurities settle 
at the bottom and are discharged through the mud blow-off. Such 
impurities as rise to the surface are removed by the surface blow-off. 
The tubes are corrugated to allow for expansion and at the same time 
to increase the transmission of heat. Referring to Fig. 352: Exhaust 
steam enters at A and leaves at E y and the portion which is condensed 
is drawn off at Z). Feed water enters at I and is discharged at 0. P, 
P are mud blow-offs and S is an opening for a safety valve. Fig. 356 
gives results of tests showing the relative efficiencies of plain and corru- 
gated tubes for various velocities. 

Fig. 353 shows a partial section through a Harrisburg feed-water 
heater. This apparatus is a typical example of the coiled-tube heater. 
Three sets of concentric copper coils are brazed to gun-metal manifolds 
and supported by clamp stays as indicated in the illustration. Feed 
water enters the heater at the bottom manifold and passes through 
the coils to the feed outlet. The exhaust steam enters the heater at 
the bottom and surrounds the coils in its passage to the outlet at the 
top. The coils are designed to withstand a pressure of 600 pounds 
per square inch. 

292. Steam-tube, Closed Heaters. — Fig. 354 shows a section through 
an Otis heater, illustrating the steam-tube type. Here the exhaust 
steam passes through the tubes which are surrounded by the feed water. 
The exhaust steam enters at A, and passes down one section of tubes 
into the enlarged space of the water and oil separator 0, in which the 
condensation and oil are deposited. From this chamber the steam 
passes up through the other section of tubes to outlet C, thus passing 
twice through the entire length of the heater. The water enters at E 
and is discharged at G. R is the blow-off opening. The tubes are of 



FEED-WATER PURIFIERS AND HEATERS 



563 



seamless brass and are curved to allow for expansion. Condensed 
steam is withdrawn at P. 

Fig. 355 shows a partial section through a Baragwanath steam 
jacketed steam-tube heater. Exhaust steam enters at A, passes up 
through the tubes, returns down annular space E between the inner 






Fig. 354. 



Otis Steam-tube Feed-water 
Heater. 



Fig. 355. Baragwanath Steam-jacketed 
Feed-water Heater. 



shell and jacket, and passes out at B. Feed water enters at C and 
leaves at D. E is the scum blow-off, G the heater drain, and H the 
jacket drain. 

293. Heating Surface, Closed Heaters. — It is generally assumed 
that the transfer of heat between two bodies is directly proportional 
to the difference in temperature between them. 



564 STEAM POWER PLANT ENGINEERING 

Let t = temperature of the water entering the heater; 

t 2 = temperature of the water leaving the heater; 

t a = temperature of the exhaust steam ; 

A = square feet of transmitting surface; 

t = temperature of a unit of water t' seconds after entering 
the heater; 

h = B.t.u. absorbed per square foot per second per degree 
difference in temperature between the steam tempera- 
ture t 8 and the water temperature t; 

t' = time in seconds; 

w = number of pounds of feed water per second; 

Then — = square feet of surface brought in contact with one Dound 
w 

of water per second; 

and dt, the rate at which the temperature of the water is increasing 
at this instant, will be 



hA 
dt = — (t 3 - t) dt'. 

w 


(201) 


dt hA j 4 , 

at . 

ts — t W 

Integrating, 

f f dt = hA fdf. 

J t s — t w J 


(202) 
(203) 


f\ dt = hA f l ' dt '. 

J to t s — t w J 


(204) 


, t s — t hAt' 
log*, , = 
t 8 — t 2 w 


(205) 


Let W = number of pounds of feed water heated per hour. 

U = B.t.u. transmitted to the feed water per square foot of surface 


per hour per degree difference in temperature. 




Then (205) may be written 




. ts-to AU 

l0ge ts-t 2 = W' 

from which 

A W, ts-to 
A =U l0 ^t 3 -t 2 ' 


(206) 
(207) 



Knowing the weight of water to be heated, the temperature of the 
steam, the desired temperature of the feed water, and the coefficient of 
heat transmission, U, this equation enables one to determine the area 
of heating surface required for the given conditions. Since the extent 
of heating surface increases rapidly as h approaches t 8 , and becomes 



FEED-WATER PURIFIERS AND HEATERS 565 

infinity for t 2 = t 8 , it is desirable to limit t 2 to some practical figure. 
An average maximum for t 2 = t a — 4. 

Table 93 has been calculated from this formula and gives the square 
feet of heating surface necessary to heat 1000 pounds of water per hour 
for different ranges in temperature. 

Mean Temperature Difference. 

If we let d = average temperature difference between the steam and 
feed water, then 

A Ud = heat given out by the steam per hour, 
W (U — to) = heat absorbed by the feed water per hour, 
AUd = W(t 2 -t ), ' ' (208) 

*=^f^- .(209) 

From (206), ^ = l oge ^|. 

Therefore d = fe ~ U • (210) 

, t s Co 

Equation 210 may be expressed 

t — t r 
d=- y (211) 

log,-, 

in which t is the original temperature difference and t' the final tempera- 
ture difference of the two fluids. Equation 211 is applicable to all 
conditions of parallel and counterflow. 

Table 94 has been calculated from formula (210) and gives the mean 
temperature difference for various conditions of operation. 

The arithmetic mean temperature difference di may be taken with 
safety for the average heater problem and has the advantage of sim- 
plicity. 

* - «. - ^ • (212) 

Closed heaters are sometimes rated on the basis of J square foot of 
heating surface per horse power, i.e., a heater with 500 square feet of 
heating surface would be rated at 1500 horse power. 

294. Heat Transmission in Closed Heaters. — An inspection of the 
curves in Figs. 320 and 356 show that the absorption of heat per square 
foot of surface per degree difference in temperature varies with the 



566 



STEAM POWER PLANT ENGINEERING 



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FEED-WATER PURIFIERS AND HEATERS 



567 



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568 



STEAM POWER PLANT ENGINEERING 



velocity of the water and the material and character of the tubes. In- 
creasing the velocity of the water passing through the heater increases 
the rate of heat transmission and thereby renders the heating surface 

more effective. In order to employ 
moderately high velocities and at 
the same time allow sufficient time 
in which to raise the temperature 
to a maximum, the tubes should be 
as long as practicable and of small 
diameters. Other things being equal, 
a heater containing a large number 
of tubes of small diameter is more 
economical than one containing a 
small number of large tubes. It is 
important to proportion the^ heater 
according to the amount of water to 
be heated and the maximum tem- 
perature to which the water must 
be raised. In designing a heater, 
then, the maximum amount of heat 
to be transmitted per degree differ- 
ence in temperature per hour per 
square foot should be assumed, and 
the velocity of the water made such 
that it is capable of absorbing this 
amount. A good average figure for 
multi-flow heaters is U = 250 B.t.u. 
for plain brass or copper tubes and 
U = 300 B.t.u. for corrugated tubes 
with a water velocity of 50 feet per minute; for single-flow heaters, 
U = 175 (for plain brass) with a water velocity of 12.5 feet per minute 
and for coil heaters U = 300 (copper) with a water velocity of 150 feet 
per minute. These figures are for water-tube heaters only. For 
steam-tube heaters (iron tubes) a good average figure is U = 120. 
(See also Fig. 356.) 

Experiments show that heaters and condensers operating with counter 
currents are more efficient and are capable of obtaining a higher final 
temperature than those operating with parallel currents. For a mathe- 
matical discussion of the parallel and counter-current flow see " Evapo- 
rating, Condensing and Cooling Apparatus/' by E. Hausbrand, Chapter I. 
It is generally supposed that the pressure of the steam and water in 
a closed heater has little influence on the heat transmission other than 




50 75 10Q 
Fig. 356. 






FEED-WATER PURIFIERS AND HEATERS 



569 



that due to increased temperature difference, but experiments recently 
conducted seem to show that the pressure has a marked influence. 
Tests made by O. M. Row (Industrial Engineering, April, 1912, p. 314) 
seem to show that the transmission of heat from steam to water falls off 
as the pressure is increased, and that when the water and steam pressures 
become equal a further increase of water pressure has no effect. Similar 
results were obtained from experiments conducted by A. H. Tuells 
(Engineering, Feb. 23, 1912). The following results were obtained from 
the tests made by Row. 

GALLONS (IMPERIAL) OF WATER PER HOUR RAISED 100° F. IN 

TEMPERATURE UNDER DIFFERENT STEAM AND 

WATER PRESSURES. 



Water Pressure, 
Lbs. per Sq. In. 


Steam Pressure, Lbs. Per Sq. In. 


100 


50 


25 


5 
10 
20 
30 
40 
50 
60 
80 
100 


215 

180 

143 

120 

104 

93 

87 

83 

82 


125 
104 
78 
64 
56 
54 
54 
54 
54 


46 
40 
36 
35 
35 
35 
35 



Transmission of Heat from Steam Through Surfaces: Engng., Feb. 9, 1912, p. 191. 

Example: Determine the size of vacuum and atmospheric heaters 
for a condensing plant of 1200 i.h.p. Engines use 20 pounds of steam 
per i.h.p. hour; auxiliaries use the equivalent of 10 per cent of the main 
engine steam; vacuum 25 inches referred to 30-inch barometer; feed 
water, t = 50 degrees; temperature of hot well, t 2 = 110 degrees; 
coefficient of heat transmission, U = 300 B.t.u. 



Vacuum or Primary Heater. 
Feed water for main engines, 

20 X 1200 = 24,000 pounds per hour. 
Feed water used by auxiliaries, 

10 per cent of 24,000 = 2400 pounds per hour. 
Total feed, 

W = 24,000 + 2400 = 26,400 pounds per hour. 



570 STEAM POWER PLANT ENGINEERING 

From formula (207), 

A W. ts-to 

A = u l0 ^t^l 2 

^ 26,400 134 - 50 

300 ge 134 -110 
= 110 square feet. 

On the basis of J square foot of surface per horse power the rating of 
this heater will be 

110 X 3 = 330 horse power. 

Atmospheric or Secondary Heater. 

The temperature of the feed water leaving the atmospheric heater, 
formula (197), will be 

. *o + 0.9 S (H + 32) 



1 + 0.9S 

1150 B.t.u., 



where 
whence 

The reqi 

where 
whence 

The hon 


S = 0.10, t = HO degrees, H = 
110 + 0.9 X 0.10 (1150 + 32) 


1 + 0.9 X 0.10 
= 198 degrees. 

lired surface is 

A W } t s -to 

A = u l °zu s -t 2 > 

t 8 = 212, *> = 110, t 2 = 198 
26,400 212 - 110 
300 ge 212 - 198 
= 175 square feet. 

5e-power rating will be 

175 X 3 = 525 horse power. 



295. Open vs. Closed Heaters. — Open and closed heaters have their 
respective advantages and a careful study of the various influencing 
conditions is necessary for an intelligent choice. The following parallel 
comparison brings out a few of the distinguishing features: 

Open Heater. Closed Heater. 

Efficiency. 

With sufficient exhaust steam for heat- The maximum temperature of the feed 
ing, the feed water may reach the water will always be 2 degrees or more 
same temperature as the steam. lower than the temperature of the 

Scale and oil do not affect the heat steam. 

transmission. Scale and oil deposit on the tubes and 

the heat transmission is lowered. 



FEED-WATER PURIFIERS AND HEATERS 571 

Open Heater. Closed Heater. 

Pressures. 

It is not ordinarily subjected to much The water pressure is slightly greater 
more than atmospheric pressure. than that in the boiler when placed 

on the pressure side of the pump as 
is customary. 

Safety. 

Sticking of the back pressure valve may It will safely withstand any pressure 
cause it to " blow up " if provision is likely to occur, 
not made for such an emergency. 

Purification. 

Since the exhaust steam and feed water Oil does not come in contact with the 

mingle, provision must be made for feed water. 

removing the oil from the steam. Scale is removed with difficulty. 

Scale and other impurities precipitated 

in the heater are readily removed. 

Location. 

Must always be placed above the pump May be placed anywhere on the pres- 
suction and on the suction side. sure side of the pump. 

Pumps. 

With supply under suction two pumps One cold-water pump is necessary, 
are necessary and one must handle 
hot water. 

Adaptability. 

Particularly adaptable for heating All vacuum or primary heaters are 
systems where it is desired to pipe necessarily of this type, 
the " returns " direct to heater. 

296. " Through " Heaters. — Fig. 357 shows a typical installation of 
a through heater in a non-condensing plant. 

It is evident that all the steam must pass through the heater. Now, 
one pound of exhaust steam in condensing gives up approximately 
1000 B.t.u. Hence, if the initial temperature of the feed water is 

50 degrees and the final temperature 210, the engine furnishes ^-tt — 

ZLO — oU 

= 6.26, say, six times the quantity necessary for heating the feed water 

to a maximum. Therefore the area of the pipe supplying the heater 

with steam need be but one sixth that of the main exhaust. With the 

heater connected as in Fig. 357 the connections must necessarily be the 

same size as the exhaust pipe. 

With this arrangement the heater cannot be "cut out" while the 

engine is in operation and hence it is not adapted for, plants working 



>72 



STEAM POWER PLANT ENGINEERING 



continuously. For the purpose of cutting out a heater while the plant 
is in operation a through heater may be by-passed as in Fig. 358. Ad- 
vantage may be taken here of the permissible reduction in the size of 
pipes and fittings, i.e., valves, etc., at C and D need be but one half 




CXHAUST TO. HEATER 

Fig. 357. Open Heater connected as a "Through" Heater. Non-condensing Plant. 

the size of those at A. This reduction in size may prove to be a con- 
siderable item in large installations. 
297. Induced Heaters. — Fig. 359 shows a typical installation of an 
induced heater in a non-condensing plant and Fig. 
360 an induced primary heater in a condensing 
plant. 

In the arrangement in Fig. 359 the number of 
fittings is reduced to a minimum and the heater 
may be readily cut out. Since induced heaters 
are apt to become air-bound, a vapor pipe or 
vent is inserted in the top of the heater as shown. 
This pipe varies from \ to 1^ inches in diameter, 
depending upon the size of heater. 

Closed Heaters: Am. Elecn., May, 1900, p. 236, July, 
1900, p. 354, Oct., 1905, p. 530; Cassier's Mag., Aug., 1903, 
p. 330; Eng. U. S., Jan. 1, 1906, p. 13; Power, April, 1902, 
p. 11. 



A(<g 




Fig. 358. 



298. Live-steam Heaters and Purifiers. — The function of a live- 
steam heater and purifier is primarily that of purification and hence it 
is not ordinarily installed unless the feed water contains scale-forming 
elements such as sulphates of lime and magnesia. These, as previously 
stated, are not entirely precipitated until a temperature of approximately 



FEED-WATER PURIFIERS AND HEATERS 



573 



300 degrees F. is reached; hence no amount of heating with exhaust 
steam at atmospheric pressure will thoroughly purify feed water con- 
taining these elements. 



Cold"Water Supply , 




Fig. 359. Open Heater Connected as an "Induced" Heater. Non-condensing Plant. 

Fig. 363 shows a section through a Hoppes live-steam purifier. Since 
the purifier is subjected to full boiler pressure, the shell and heads are 
constructed of steel. Within the shell are a number of trough-shaped 



To Atmos 




Fig. 360. Closed Heater Connected as an "Induced" Heater. Condensing Plant. 

pans or trays placed one above another and supported on steel angle 
ways. Steam from the boiler enters the chamber at A and comes in 
contact with feed water and condenses. The water on entering the 



574 



STEAM POWER PLANT ENGINEERING 



Back Pressure Valve 



-Noa.Return.Air Valves 




Vacuum Pump 



Fig. 361. Open Heater in Connection with a Low-pressure Turbine. 



To Atmosphere 

Back Pressure 
Valve 



Ezh.from Fan 
and Stoker [ 

Engs.Pumps. 5 1 

Etc. 



. — ■ — — : — 




Outflow Intake 

Fig. 362. Open Heater in Connection with a Jet Condenser. 



FEED-WATER PURIFIERS AND HEATERS 



575 



heater at B is fed into the top pan and, overflowing the edges, follows 
the under side of the pan to the center and drops into the pan below. 
It flows over each successive pan in the same manner until it reaches 




Fig. 363. Hoppes Live-steam Purifier. 

the chamber at the bottom, whence it gravitates to the boiler through 
pipe C. As the steam inclosed in the shell comes in contact with the 
thin film of water, the solids held in solution are separated and adhere 




Fig. 364. Typical Installation of a "Live-steam" Purifier. 

to the bottom of the pans in the same manner that stalactites form on 
the roofs of natural caves. Authentic tests show that live-steam heaters 
may increase the boiler efficiency. (See Power, Feb. 21, 1911, p. 295.) 
The purifier should be set in such a position as will bring the bottom of 



576 



STEAM POWER PLANT ENGINEERING 



the shell two feet or more above the water level of the boilers, as in Fig. 
364. N is the feed pipe from pump to purifier and should be provided 
with a check valve. D is the gravity pipe through which the purified 
water flows to the boiler. This pipe should be carried below the water 
level of the boilers and all branch pipes should be taken off below the 
water line. Pipe L leads from top of pipe S to pump or other steam- 
using device. This is necessary in order that air and other non-condens- 
able gases liberated from the water may be removed from the purifier, 
which would otherwise become air-bound. In the illustration the feed 
pump takes its supply from an exhaust steam heater C The purifier 
is provided with a suitable by-pass so that the water may be fed directly 
to the boiler when necessary. 

Live Steam Heated Feed Water: Elec. Engr., Lond., June 29, 1906; Cassier's Mag., 
Oct., 1911, p. 543; Elec. Rev., Lond., May 20, 1898, p. 667; Eng. Ree., Aug. 30, 
1898, p. 467; Power, March 31, 1908, p. 498, Feb. 21, 1911, p. 295. 




Fig. 365. Green Economizer. 



299. Economizers. — Fig. 365 gives a general view of a Green econo- 
mizer, illustrating a typical flue gas heater. It consists of a series of 
cast-iron tubes 9 to 10 feet in length and 4| inches in diameter, which 
are arranged vertically in sections of various widths across the main 
flue between boiler and chimney. When in position the sections are 
connected by top and bottom headers, and the headers are connected 
to branch pipes running lengthwise, one at the top and the other at 
the bottom. Both of the branch pipes are outside the brickwork which 
incloses the apparatus. The waste gases are led to the economizer by 
the ordinary flue from the boiler to the chimney, but a by-pass must be 
provided for use when the economizer is out of service for cleaning or 
for repairs. The feed water is forced into the economizer through the 



FEED-WATER PURIFIERS AND HEATERS 



577 



lower branch pipe nearest the point of exit of gases, and emerges through 
the upper branch pipe nearest the point where the gases enter. Each tube 
is encircled with a set of triple overlapping scrapers which travel con- 
tinuously up and down the tubes at a slow rate of speed, the object 
being to keep the external surfaces free from soot. Tl^e mechanism 
for working the scrapers is placed on top of the economizer, outside 
the chamber, and the motive power is supplied either by a belt from 
some convenient shaft or small independent engine or motor. The 
power for operating the gearing varies from 1 to J horse power per 
1000 square feet of economizer surface, depending upon the number 
and length of tubes. The apparatus is fitted with blow-off and safety 
valves, and a space is provided at the bottom of the chamber for the 



5 Saturated Header 




Fig. 366. Typical Economizer Installation. 



collection of soot. For continuous plant operation the soot is automati- 
cally cleaned as shown in the illustration. 

This type of economizer is also used as an air heater for drying and 
heating purposes. The air heater is similar in design to the water heater 
with the exception of the direction of flow and size of tubes. The tubes 
in the air economizer are 3| inches internal diameter by 9 feet in length, 
as against 4f inches internal diameter for the water economizer. In 
the latter the water enters at the bottom header and passes out from 
the top header, in the former the air is forced by a fan first through one 
set of tubes and up through another set, and then down again, and so 
on until it leaves the heater. 

300. Value of Economizers. — The general conclusion drawn from 
current practice is that an economizer installation results in: 

A small annual saving in cost of operating the plant. 



578 STEAM POWER PLANT ENGINEERING 

Decreased wear and tear on the boilers due to the higher feed-water 
temperature. 

A large storage of hot water for sudden increase in load. 

Purification of the feed water due to the high temperature in the econo- 
mizer. The scale-forming elements do not bake hard on the economizer 
tubes as they would in the boiler where the heat from the fire is more 
intense, but make a muddy deposit readily removed by blowing off. 

301. Factors Determining Installation of Economizers. — The factors 
to be considered before installing an economizer are: 

The nature of the auxiliary machinery, direct connected or belted. 

Method of heating the feed water; whether vacuum and atmos- 
pheric heaters are used and whether all or part of the auxiliary steam 
is used for heating. 

Initial temperature of the feed water; whether the feed is taken 
from the hot well or from a cold supply. 

Rise in temperature due to economizer. 

Cost of economizer. An approximate price is $15 per tube erected, 
on a basis of 15 square feet per tube. The heating surface is rated at 
3 to 5 square feet per boiler horse power. 

Cost of additional building space. 

Reduction in boiler-heating surface made possible by the economizer. 

Extra cost of stack or forced-draft apparatus necessary to compensate 
for loss of draft due to economizer. The economizer lessens the draft 
by increasing the resistance between boilers and chimney and by re- 
ducing the chimney temperature. Where the installation of an econo- 
mizer decreases the normal temperature of the chimney from, say, 
550 degrees to 350 degrees F., the reduction in draft is approximately 
25 per cent. 

Total cost of economizer plant. This depends largely upon the 
design and varies from $4 to $7 per boiler horse power. 

Interest, depreciation, repairs, operation, taxes, and insurance. 

Table 95 gives the results of economizer tests. 

302. Temperature due to Use of Economizer. — The rise in tem- 
perature of feed water due to the use of an economizer may be approxi- 
mated from the following empirical formula advocated. by the Green 
Economizer Company: 

v M - *■) - 

o i ■ 5^ + QC (213) 

9 - 1+ 2GC y 
in which 

x = rise in temperature of the feed water, 

T\ = temperature of flue gas entering economizer, 



FEED-WATER PURIFIERS AND HEATERS 



579 



h = temperature of feed water entering economizer, 
w = pounds of feed water per boiler horse power per hour, 
G = pounds of flue gas per pound of combustible, 
C = pounds of coal per boiler horse power per hour, 
y = square feet of economizer heating surface per boiler horse 
power. 




Fig. 367 



Referring to Fig. 367, the ordinates represent temperatures, and 
abscissas the path of the flue gas and the water in the economizer. 
The flue gas enters the economizer at c with temperature Ti and leaves 
at a with temperature T. The feed water enters at b with temperature 
t\ and leaves at d with temperature h + x. 

The algebraic mean temperature difference D between the flue gas 
and the feed water will be 

cd + ab 



D = 



T x 


- (h + x) + of- 


- oa — 


bf 


T, 


2 

- fa + x ) + T u - 


- oa — 


h 



= Ti-*i 



x + oa 



(214) 



(215) 



(216) 



(217) 



Now, wx = B.t.u. absorbed by the feed water per boiler horse power 
and 
GCS = B.t.u. given up to the feed water by the flue gas for 
each degree reduction in temperature (S = mean 
specific heat of the flue gas); therefore 



* The theoretical mean temperature difference is d = 



ti -U 

i *i 

logei 

h 



in which t\ = 



original temperature difference and U = final temperature difference between the 
two fluids. 



580 STEAM POWER PLANT ENGINEERING 

wx + GCS = total reduction in temperature of the flue gas, that is, 

wx 
GCS = ° a ' < 218 > 

Substituting (218) in (217), we get 

. wx 
x-\- 



D = Tl - k - —™* W 

in which 

D — mean temperature difference between flue gas and feed 
water, degrees F. 
Let U = B.t.u. absorbed per hour per square foot per degree 
difference in temperature and 
y = square feet of economizer surface per boiler horse power. 
Then U Dy = heat absorbed per boiler horse-power hour. 
But wx = heat absorbed per boiler horse-power hour. 
Therefore wx = U Dy. (221) 

Combining (220) and (221), 

tt fm , xGCS + wx\ , . 

wx = UylTx -h 2 QCS J, (222) 



from which x ~ 



w w + GCS (223) 

*7 + 2 GCS 



y varies from 3.5 square feet to 5 square feet per boiler horse power, 
and U from 2.25 to 3.3, depending upon the conditions of operation.* 

If we let w = 30, S = 0.2, and U = 3.3, and substitute these values 
in equation (223), it assumes the form given by the Green Economizer 
Company, equation (213). 

A method of approximating the rise in temperature where the final 
temperature of the flue gas is known, is to assume T % degree rise in the 
feed water for each degree reduction in temperature in the flue gas. 
This is determined on the basis that approximately 20 pounds of flue 
gas are generated for each pound of combustible, and that 10 pounds 
of water are evaporated per pound of combustible; that is, 2 pounds 
of flue gas are generated for each pound of feed water delivered to the 
boiler. Assuming a specific heat of 0.25 for the flue gas, this gives 
2 X 0.25, or 0.5 degree rise in temperature in the feed water for each 
degree reduction in the flue gas temperature. 

* For D = 600 U = 3.25 For D = 400 U = 2.75 

500 3.00 300 * 2.25 



I 



FEED-WATER PURIFIERS AND HEATERS 



581 



Example: Determine the rise in temperature of the feed water in a 
power plant of 1200 i.h.p. Engines use 20 pounds of steam per i.h.p. 
hour; auxiliaries use the equivalent of 12 per cent of main engine 
steam; vacuum 25 inches; feed- water supply 50 degrees; 3.7 pounds 
of coal are burned per hour per boiler horse power; flue gas temperature 
550 degrees F.; steam pressure 150 pounds gauge. 

The vacuum and atmospheric heater will raise the temperature of the 
feed water from 50 to 205 degrees. (See preceding problem.) 

On the assumption that 20 pounds of flue gas are generated per 
pound of combustible and that 3.5 square feet of economizer heating 
surface are installed per boiler horse power, the notations in the formula 
will become T x = 550, k = 205, w = 30, G = 20, C = 3.7, y = 3.5, 
U = 3.3, S = 0.2. 

Substituting, 3.5 (550 - 205) 

X 30 (5X30 + 20X3.7) 
3.3 + (2 X 20 X 3.7) 
= 83 degrees rise in temperature. 

Therefore the temperature of the water entering the boiler will be 

205 + 83 = 288 degrees F. 

Economizers: Prac. Engr. U. S., March, 1910, May, 1910, p. 282, July 15, 1912, 
Power, July 27, 1909; Cassier's, March, 1900, p. 378; Eng. Mag., June, 1912, 



736; 
389. 



TABLE 95. 
ECONOMIZER PERFORMANCES. 



Number of 
Plant. 



Number of 
Economizer 

Tubes 
Installed. 



Temperatures, Deg. Fah. 



Gases 

Entering 

Economizer 



Leaving 
Economizer 



Fluid 

Entering 

Economizer 



Fluid 

Leaving 

Economizer 



Rise in 

Temperature 

of Fluid. 



Actual 
Saving in 

Fuel, 
Per Cent. 



Water Heater. 



1 


160 


435 


279 


84.2 


196.2 


112.0 


12.5 


2 




416 
620 


254 
293 


40.0 
101.0 


185.4 
237.0 


125.4 
136.0 


13.8 


3 


960 


18.3 


4 


520 


548 


295 


96.0 


200.0 


104.0 


9.2 


5 


520 


603 


325 


93.5 


203.8 


110.3 


9.7 


6 


384 


368 


245 


103.0 


202.6 


99.6 


12.4 


7 


448 


537 


326 


71.2 


203.4 


132.2 


17.5 



Air Heater. 



72 


301 


257 


70.0 


152.0 


240 


512 


319 


54.0 


201.6 


96 


557 


376 


41.0 


200.0 


192 


417 


369 


74.0 


210.0 



82.0 
147.6 
159.0 
136.0 



9.0 
14.0 



Compiled from " The Book of the Economizer," 1912, published by the Green Engineering Co. 



582 STEAM POWER PLANT ENGINEERING 

303. Choice of Feed-water Heating System. — The heating of feed 
water and its delivery to the boiler in the most economical manner 
is a problem involving such a large number of combinations that a 
general analysis is impracticable. The following discussion of a specific 
case will give some idea of the manner in which this problem may be 
attacked. 

Example: Determine the most economical manner of heating the 
feed water for a power plant of 1000 horse power operating under the 
following conditions: Schedule 10 hours per day and 310 days per 
year; load factor on the ten-hour basis 0.8; cost of coal $2.50 per 
ton of 2000 pounds; heat value of the coal 13,500 B.t.u. per pound; 
average boiler efficiency 65 per cent; engines use 20 pounds of steam 
per i.h.p. hour; steam pressure 150 pounds absolute; temperature of 
cold water 60 degrees; vacuum 26 inches referred to 30-inch barometer; 
interest 5 per cent; depreciation 8 J per cent; maintenance 1 per cent; 
insurance \ per cent; taxes 1 per cent; total charges 16 per cent; 
charges for attendance and maintenance assumed to be the same in 
each case and credit for the chimney assumed to offset debit for econo- 
mizer space. Many of the influencing conditions are left out for the 
sake of simplicity. 

The most likely combinations are 

(1) Atmospheric, all auxiliaries steam driven, water taken from cold 

well. 

(2) Same as (1) except that water is taken from hot well. 

(3) Economizers, auxiliaries electrically driven, chimney draft, water 

from cold well. 

(4) Vacuum heater, economizer, and electrically driven auxiliaries, 

fan draft. 

(5) Vacuum heater, atmospheric heater, and steam auxiliaries. 

(6) Atmospheric heater, economizer, steam auxiliaries, fan draft. 

(7) Vacuum and atmospheric heaters, economizers, steam auxiliaries, 

and electrical fan. 

(8) Vacuum, atmospheric heater, economizer, and chimney draft, 

auxiliaries operating condensing except feed pumps and stoker 
engines which exhaust into the atmospheric heater. 

The difference between the total heat furnished by the boiler and 
the heat returned in the feed water is the net heat put into the steam 
by the boiler. Evidently the system which shows the least net heat 
required to produce one horse power will be the most economical as 



FEED-WATER PURIFIERS AND HEATERS 583 

far as coal consumption is concerned, although not necessarily the 
cheapest when both operating and fixed charges are considered. 

Prices vary so much that it is practically impossible to give costs of 
installations which will bear criticism and the prices taken in this problem 
are approximate only. 

Case I. 

Atmospheric heater, auxiliaries steam driven, feed from cold well. 

This arrangement and that of Case II are the most common in power 
plants of this size. 

The power consumption of the auxiliaries operating non-condensing 
varies from 8 to 12 per cent of the total power developed. Assume 
it to be 10 per cent. 

The temperature of the feed water leaving the heater may be deter- 
mined by formula (197). 

t Q + 0.9 S (X + 32) 
1 + 0.9 S 

Substituting S = 0.10, X = 1146, t = 60, 

_ 60 + 0.9 X 0.10 (1146 + 32) 

1 + 0.9 X 0.10 
= 152. 

The net heat furnished by the boiler to produce one indicated horse- 
power hour in the engine is evidently the heat necessary to raise 20 + 10 
per cent of 20 = 22 pounds of water from 152 degrees F. to steam at 
150 pounds pressure; i.e., the net heat furnished is 

22 X 1071.2 = 23,564 B.t.u. 

Now, 1 i.h.p. = 2546 B.t.u. 

Therefore the heat efficiency of this arrangement is 

2546 1AQ 

= 10.8 per cent. 



23,564 

Probable First Cost. 

Steam pumps $ 400.00 

Condenser with steam-driven air and circulating pumps 3000.00 

1000-horse-power open heater 480.00 

Piping 1200.00 



$5080.00 



584 STEAM POWER PLANT ENGINEERING 



Fuel Consumption. 

Average horse-power hours per year = 1000 (rated horse power) X 0.8 (curve 
load factor) X 310 (days per year) X 10 (hours per day) = 2,480,000. 

Pounds of coal per i.h.p. hour = net heat furnished per i.h.p. hour -r net heat 
absorbed by the boiler per pound of coal = 23,564 -r- (13,500 X 0.65) = 2.68. 

rp 2,480.000 X 2.68 QQOO 
Tons per year = - -^oqo = 3323. 



Fuel and Fixed Charges. 

Fuel, 3323 tons at $2.50 $8308.00 

Fixed charges, 16 per cent of $5080 812.00 



$9120.00 
Case II. 

Same as Case I, except that feed is taken from the hot well. This 
arrangement is possible only when the condensing water is suitable 
for feed purposes. 

Assume the temperature of the water from the hot well as it enters 
the heater to be 110 degrees. 

The temperature of the feed water leaving the heater will then be 
198 degrees (from formula (197)). 

Net heat furnished = 22 X 1025.2 = 22,554 B.t.u. 

Efficiency = 99 . = 11.3 per cent. 

22 554 
Pounds of coal per i.h.p. hr. = ^ x Q 65 = 2.62. 

„ 2,480,000 X 2.62 QO 
Tons per year = -? ^joo = 3248. 



Fuel and Fixed Charges. 

Fuel, 3248 tons at $2.50 $8120.00 

Fixed charges (same as Case I) 812.00 



$8932.00 
Case III. 

Economizers, auxiliaries electrically driven, chimney draft, water 
from the cold well. 

Practice gives an average of 3 per cent of the main engine output as 
the power required to operate the electrical auxiliaries in a plant of this 
size. 



FEED-WATER PURIFIERS AND HEATERS 585 

The temperature of the feed water leaving the economizer may be 
determined from formula (213). 



9A + -2G^~ y 



Substituting, 



3.5 (550 - 60) ,,„ , 

x = n , , 5x30+20x3 -7 7: = 119 degrees - 

9,1 + 2X20X3.7 6 - b 

Temperature of feed water entering heater = 119 + 60 = 179 degrees. 
Net heat furnished = (20 + 3 per cent of 20) X 1044.2 = 21,510 
B.t.u. 

2545 
Efficiency = =11.8 per cent. 

^l,olU 

Probable First Cost. 

Economizers $3500.00 

Motor feed pump 600.00 

Condenser with electrically driven air and circulating pump .... 6000.00 

Piping and wiring 1000.00 



$11,100.00 
Fuel Consumption. 

Pounds of coal per i.h.p. hour = ^ ^^ = 2.45. 

™ 2,480,000 X 2.45 _„. 
Tons per year = -* ^qoo = 3038 " 

Fuel and Fixed Charges. 

Fuel, 3038 tons at $2.50 $7595.00 

Fixed charges, 16 per cent on $11,100 1776.00 

$9371.00 

Case IV. 

Vacuum heater, economizer, electrically driven auxiliaries, fan draft. 

The vacuum heater may be relied upon to raise the temperature of 
the feed water to 110 degrees. 

The economizer will increase this 107 degrees (from formula (213)), 
giving the feed water a temperature of 217 degrees as it enters the 
heater. 

The electrical fan for the mechanical-draft system will require ap- 
proximately 2 per cent of the main system engine power, making a 
total of 3 + 2 = 5 per cent for all auxiliaries. 



586 



STEAM POWER PLANT ENGINEERING 



Net heat furnished = (20 + 5 per cent of 20) X 1006.2 
= 21,130 B.t.u. 
2545 



Efficiency 



21,130 



12.05 per cent. 



Probable First Cost. 

For the sake of simplicity it is assumed that the high first cost of the 
chimney plus its low depreciation and maintenance will offset the low 
first cost of the mechanical-draft system plus its higher maintenance 
and depreciation charges: 

Economizers $3500.00 

Motor feed pump 600.00 

Motor-driven pumps and condenser . 6000.00 

Motor-driven fan 750.00 

Piping and wiring 1200.00 

Vacuum heater 200.00 

$12,250.00 
Fuel Consumption. 

Pounds of coal per i.p.h. hour = ' =2.41. 

lOjOUU X O.00 

_ 2,480,000 X 2.41 OQQQ 
Tons per year = - ^^oocj = 2988 " 

Fuel and Fixed Charges. 

Fuel, 2988 tons at $2.50 $7470.00 

Fixed charges, 16 per cent of $12,250 1960.00 

$9430.00 

In like manner Cases V, VI, VII, and VIII have been treated and are 
tabulated in the summaries. 



SUMMARY (1). 



Case. 


Temperature 
of Feed 
Water. 


Power 
Consumed by- 
Auxiliaries. 


Efficiency. 


First 
Cost. 


Fuel Cost 
per Year. 


Cost of 
Operation 
per Year. 


I 


Degrees F. 
152 
198 
179 
217 
208 
294 
290 
270 


Per Cent. 
10 
10 

3 

5 

10 
14 
10 

8 


Per Cent. 
10.8 
11.3 
11.8 
12.05 
11.4 
12 

12.2 
12.3 


$5,080 
5,080 
11,100 
12,250 
5,280 
9,000 
9,300 
8,250 


$8,308 
8,120 
7,595 
7,470 
7,900 
7,750 
7,380 
7,075 


$9,120 


II 


8,932 


Ill 


9,371 


IV 


9,430 


V 


8,744 


VI 


9,190 


VII 

VIII 


9,570 
8,395 



FEED-WATER PURIFIERS AND HEATERS 



587 



SUMMARY (2). 



Case. 


Efficiency. 


First Cost. 


Fuel. 


Cost per Year. 


I 


8 
7 
6 
3 
5 
4 
2 
1 


1 
1 
6 
7 
2 
4 
5 
3 


8 
7 
4 
3 
6 
5 
2 
1 


4 


II 


2 


Ill 


6 


IV 


7 


V 


3 


VI 


5 


VII 


8 


VIII 


1 



Summary (2) gives the ranking; thus: Case I is eighth in point of 
efficiency; first in cheapness of installation; eighth in yearly cost of 
fuel; and fourth in yearly cost of operation. Case VIII is apparently 
the best arrangement for the given conditions. 



CHAPTER XIII. 

PUMPS. 

304. Classification. — Pumps used in connection with steam power 
plants may be conveniently classified under five groups according to 
the principles of action. 

1. Piston pumps, in which motion and pressure are imparted to 
the fluid by a reciprocating piston, plunger, or bucket. The action is 
positive and a certain definite amount of fluid is handled per stroke 
under predetermined conditions of pressure and velocity. 

2. Centrifugal pumps, in which the fluid is given initial velocity and 
pressure by a rotating impeller. The action is not positive, as the 
amount of fluid discharged is not necessarily proportional to the im- 
peller displacement. 

3. Rotary pumps, in which motion and pressure are imparted to 
the fluid by a rotating impeller. The volume discharged is practically 
equal to the impeller displacement regardless of pressure. 

4. Jet pumps, in which velocity and pressure are imparted to the 
fluid by the momentum of a jet of similar or other fluid. The ordinary 
steam injector is the best known of this group. 

5. Direct-pressure pumps, in which the pressure of one fluid acts 
directly on the surface of another fluid, thereby imparting all or part 
of its energy to the latter. The pulsometer is an example of this type. 

These groups may be variously subdivided as follows: 



Piston. 



direct-acting.. { »■ 

Fly-wheel J Simplex. 

Power driven.. ^p e x. 
Triplex. 



Air. 

Vacuum. 
Forcing. 
Lifting. 



Centrifugal . . . . { £££. ; •■•••• *»$*«£ 

** i *>*»*» ! luetic- 

Direct pressure j £*~ ; ; ; ; ; ;£££§• • ■ ; 

Piston or plunger pumps are the most common in use. Boiler-feed 
pumps, city waterworks pumps, and force pumps are ordinarily of this 
type. In the direct-acting type, Fig. 369, the water plunger and 
steam piston are secured to a single piston rod and the steam pressure 

588 



PUMPS 589 

is transmitted directly to the water. There is no flywheel, connect- 
ing rod, or crank. The velocity of the delivery is proportional to the 
resistance offered by the water; when the resistance equals the forward 
effort of the steam pressure the pump stops. This class of pump is 
well adapted for boiler-feeding purposes, since it may be operated as 
slowly as suits the requirements of feeding by simply throttling the 
discharge. The steam consumption is very large in proportion to the 
work performed, since the steam is not used expansively. 

Flywheel pumps, Figs. 382, 423, are ordinarily classified as pumping 
engines. In this class steam may be used expansively, as sufficient 
energy is stored in a flywheel to permit the drop in steam pressure 
during expansion. These pumps find wide application in city water- 
works, elevator plants, and the like, where high duty is required. They 
are little used as stationary boiler feeders, but are used to some extent 
in river-boat practice and in plants operating continuously for long 
periods at comparatively steady loads. Practically all sizes of dry-air 
pumps and a number of large jet condenser pumps are of this type. 

Piston pumps, Fig. 389, driven by gearing or belting are ordinarily 
classified as power-driven pumps. The driving power may be steam 
engine, electric motor, or gas engine. The single-cylinder machine is 
often designated as a " simplex" power-driven pump, the two-cylinder 
as a " duplex," the three-cylinder as a "triplex," and so on. 

Centrifugal pumps, Fig. 409, are supplanting to a considerable" extent 
the present type of piston pump for many uses. Though particularly 
adapted for low heads and large volumes they are used in many situ- 
ations requiring extremely high heads. They are not as efficient as 
high-grade pumping engines, but the extremely low first cost fre- 
quently offsets this disadvantage, and they are much used in connection 
with dry docks, irrigating plants, sewage systems, and as circulating 
and vacuum pumps in condensing plants. 

Rotary pumps, Fig. 421, are employed to a limited extent in the 
same field as the centrifugal pump. Being positive in action, they 
permit of a much lower rotative speed for the same delivery pressure. 

Jet pumps, Fig. 393, are seldom used as pumps in the ordinary sense 
of the word, on account of their extremely low efficiency, but are fre- 
quently employed for discharging water from sumps. Their greatest 
field of application lies in boiler feeding and in this respect their effi- 
ciency is comparable with that of the average piston pump. 

Direct-pressure pumps operated by steam, such as the "pulsometer," 
Fig. 424, are used principally for pumping out sumps, surface drains, 
and the like, where the operation is intermittent. Direct-pressure 
pumps of the air-lift type, Fig. 425, are quite common and are used a 



590 



STEAM POWER PLANT ENGINEERING 



great deal in situations where water is to be pumped from a number 
of scattered wells. 

305. Boiler-feed Pumps, Direct-acting Duplex. — Figs. 368 and 369 
illustrate a typical duplex boiler-feed pump, which consists virtually of 
two direct-acting pumps mounted side by side, the water ends and the 
steam ends working in parallel between inlet and exhaust pipe. The 
piston rod of one pump operates the steam valve of the other through 
the medium of bell cranks and rocker arms. The pistons move alter- 
nately, and one or the other is always in motion, the flow of water 
being practically continuous. 



Air 

Chamber 



Discharge 




Fig. 368. Typical Duplex Pump. 

In general construction the steam pistons and valves are similar 
to those of steam engines. The valves in duplex pumps, however, 
have no lap. In order to reduce the valve travel to a minimum, and 
still have sufficient bearing surface between the steam ports and the 
main exhaust ports to prevent the leakage of steam from one to the 
other, separate exhaust ports are provided which enter the cylinder 
at nearly the same point as the steam ports. This arrangement offers 
a simple means of cushioning the piston by exhaust steam, thus pre- 
venting it from striking the cylinder heads at the ends of the stroke. 
The valves of the duplex pump having no lap would, if connected 
rigidly to the valve stem, open one port as soon as the other had been 
closed, at about mid-stroke of the piston, thus cutting down the stroke 



PUMPS 



591 



to about one fourth the usual length. To obviate this difficulty the 
valves are given considerable lost motion by allowing sufficient clear- 
ance between the lock nuts on the valve stem; the latter, therefore, 
imparts no motion to the valve until the piston operating it has nearly 



DISCHARGE 




Valve Stem 



Fig. 369. Section through a Typical Duplex Boiler-feed Pump. 

completed the stroke. The lost motion between valves and lock nuts 

renders it impossible to stop the pump in any position from which it 

cannot be started by simply admitting steam, and therefore the pump 

has no dead centers. When one piston moves to the end of the stroke 

it pulls or pushes the opposite 

valve to the end of its travel; 

then when the piston starts back 

to the other end of its stroke 

the valve remains stationary, 

owing to the lost motion, until 

the piston has completed about 

one half the stroke. During 

this time the opposite piston has Fig- 370. 

completed a full stroke and the valve operated by it will have opened 

the steam port wide, so that while one valve covers both steam ports 

the other is at the end of its travel. In some makes of pumps the stem 

is rigidly attached to the valves, the lost motion being adjusted outside 

the steam chest as shown in Figs. 370 and 371, which represent two 

common constructions of duplex valve gear. 




592 



STEAM POWER PLANT ENGINEERING 



Valve Stem 




| Piston Rod 



Fig. 372 shows the valve and piston in the position occupied at the 

commencement of the stroke. At one end of the valve the steam port 

P is open wide and at the oppo- 
site end the exhaust port E is 
open wide. When the piston 
nears the opposite end of the 
stroke and reaches the position 
shown in Fig. 373 the steam 
escape through the exhaust port 
E is cut off by the piston, and 
since the steam port is closed, 

the remaining steam is compressed between the piston and cylinder head, 

thus arresting the motion of the piston gradually without shock or jar. 
The construction of the water 

end of single-cylinder and duplex 

pumps is practically the same; 

any slight differences which may 

be found are confined to minor 

details which in no way affect 

the general design or operation of 

the pump. The piston is double 

acting, the single-acting cylinder 

being confined to power pumps 

or to steam pumps intended for 

very high pressures. In the old- FlG - 372 - 

style pumps it was the custom to use one large valve with a lift sufficient 

to give the required passage, but in modern practice the required area 

is divided among several small 
valves, so that each one is easily 
and cheaply removed in case of 
accident or wear, and slip is 
lessened.* 

The valves are carried by two 
plates or decks, the suction 
valves being attached to the 
lower plate and the delivery 
valves to the upper one, as 
shown in Fig. 369. 

The valves in practically all 

boiler-feed pumps are of the flat disk type, Fig. 374, held firmly to the 

seat by conical springs and guided by a bolt through the center. 

* The modern Riedler pump is an exception. See Engineer, U. S., Nov. 15, 1907,, 
p. 1040. 





Fig. 373. 



PUMPS 



593 



All pumps are provided with an air chamber on the discharge side, 
which acts as a cushion for the water, prevents excessive pounding, and 
insures a uniform flow. Fig. 375 shows a section through the steam 
end of a compound duplex pump. 

306. Feed Pumps with Steam-actuated 
Valves. — Single-cylinder direct-acting 
pumps, Fig. 376, are ordinarily operated 
by steam-actuated valves. The steam 
enters the chest C and passes to the left 
through* the annular opening A formed 
between the reduced neck of the valve 
and the bore of the steam chest. It is 
thus projected against the inside surface 
of the valve head H before escaping 
through the port P and passing to the 
cylinder. Both the pressure and impulse 
due to velocity acting on the valve head H tend to close or restrict the 
admission port by forcing the valve to the left. On reaching the cylinder 
and forcing the piston X toward the right, the pressure of the steam upon 




Fig. 374. A Typical Pump 
Disk Valve. 




Fig. 375. Section through Steam Cylinders of a Typical Compound Duplex Pump. 

the opposite side of the valve head H is pressing the valve to the right, a 
movement which would give the admission more port opening at A 
and deliver more steam to the cylinder. The valve then holds a position 
depending upon the relative intensity of the two pressures, which tend 
to move it in opposite directions, the admission steam, tending to close 
the valve, and cylinder steam, tending to open the valve wider. The 
steam valve, therefore, is always in a balanced position. The steam 
piston is grooved at the center, forming a reservoir for live steam R 
which is supplied from the upper chamber of the steam chest by passage 



594 



STEAM POWER PLANT ENGINEERING 



E to the cylinder cap S, and thence by tube M and the hollow piston 
rod V. The steam in this annular piston space reverses the steam valve 
by pressing alternately against the outer surfaces of the valve heads H 
through the connecting passages 0, near each end of the cylinder. 
The tappets T are for the purpose of moving the valve by hand in case 
it fails to move automatically. Steam-actuated valves are not as 
positive in action as mechanically operated valves, and hence are little 
used in situations where positive action is essential, as in fire-pump 
service, 




Fig. 376. Marsh Boiler-feed Pump. A Typical Steam-actuated Valve Gear. 

307. Air and Vacuum Chambers. — Air chambers in piston pumps 
are for the purpose of causing a steady discharge of water and of re- 
ducing excessive pounding at high speeds by providing a cushion for 
the water. The water discharged under pressure compresses the air 
in the air chamber somewhat above the normal pressure of discharge 
during each stroke of the water piston, and when the piston stops 
momentarily at the end of the stroke the air expands to a certain ex- 
tent and tends to produce a uniform rate of flow. 

The volume of the air chamber varies from 2 to 3J times the volume 
of the water piston displacement in single-cylinder pumps, and from 



PUMPS 



595 



1 to 2\ times in the duplex type. High-speed pumps are provided with 

air chambers of from 5 to 6 times the piston displacement. The water 

level in the air chamber should be kept down to one fourth the height 

of the chamber. In slow-running pumps 

sufficient air may be carried into the pump 

chamber along with the water, but with 

high speeds a large part of the air will be 

discharged, and air must be forced into the 

chamber by mechanical means. The larger 

the chamber the more uniform will be the 

discharge pressure. 

Vacuum chambers are frequently pro- 
vided for the purpose of maintaining a 
uniform flow of water in the suction pipe 
and assisting in the reduction of slip. Such 
chambers should be of slightly greater vol- 
ume than the suction pipe and of considerable length rather than diameter. 
Fig. 377 illustrates two designs commonly used. The one in Fig. 377 (B) 
should be placed in such a position as to receive the impact of the 



Fig. 377. Forms of Vacuum 
Chambers. 




Fig. 378. Different Arrangements of Vacuum Chambers. 

column of water in the suction pipe as illustrated in Fig. 378 (A), (B) 
and (C). The chamber illustrated in Fig. 377 (A) should be placed in 
the suction pipe below but close to the pump. 

308. Water Pistons and Plungers. — In cold-water pumps the water 
pistons are usually packed with some kind of soft packing. Fig. 379 (A) 
shows the details of a piston with square hydraulic packing. The body 
E is fastened to the piston rod by nut C; packing is placed at D, and 
follower F is forced up by the nut B and locked by nut A. For large 
sizes the design is the same except that the follower is set up by a 
number of nuts near the edge. In hot- water pumps the pistons are 
often packed by means of metallic piston rings R, R, Fig. 379 (C), similar 



596 



STEAM POWER PLANT ENGINEERING 



to those in steam pistons, or merely by water grooves G, G, Fig. 379 (B). 
The water end is often fitted with a plunger instead of a piston, as in 
Figs. 380 to 382. The piston is more compact, but the plungers do 
not require a bored cylinder, so that the first cost is not materially 
different. 



G G 




(B) (C) 

Fig. 379. Types of Water Pistons. 

Fig. 380 shows a plunger with metal packing ring. When leakage 
becomes excessive it is necessary to renew the ring, which is readliy 
removed. 

In Fig. 381 the plunger is packed with hydraulic packing as in the 
follower type of pump piston. The great difficulty with the above 




Fig. 380. Plunger with Metal Packing Ring. 

types of piston and plunger is in keeping the packing tight or in know- 
ing when it is leaking, and the trouble necessary to replace the packing. 
The outside packed plunger, Fig. 382, obviates these disadvantages to a 
great extent, since leakage is readily detected and repacking is performed 
without removing the cylinder heads. In dirty, dusty locations, how- 
ever, the piston pump or inside packed plunger is to be preferred, 
since the abrasive action of the dust renders outside packing difficult. 






PUMPS 



597 



Fig. 382 illustrates a high-duty elevator pump with outside packed 
plunger. 

309. Performance of Piston Pumps. — Direct-acting pumps as a 
class are wasteful of fuel and low in efficiency, due largely to the non- 
expansive use of steam. The average small duplex boiler-feed pump 
uses from 100 to 200 pounds of steam per i.h.p. hour, depending upon 
the speed, and the mechanical efficiency varies from 50 per cent to 90 
per cent. When new and in proper working condition the mechanical 
efficiency is seldom less than 85 per cent; but such pumps, as a rule, 
are given scant attention, and the average efficiency is not far from 




'XtwA kA k 




mmm 



AMJL& 



1 ; ' 



z~ 



IQ 





VM/////////////////A 



Fig. 381. Plunger with Hydraulic Packing. 



65 per cent. The term " mechanical efficiency" in this connection re- 
fers to the ratio of the actual water horse power to the indicated horse 
power of the steam cylinder. The loss includes the slip of the piston 
and valves. A steam consumption of 150 pounds per i.h.p. hour with 
mechanical efficiency of 65 per cent is equivalent to a power con- 
sumption of about 5 per cent of the rated boiler capacity, although if 
the exhaust steam is used for feed-water heating the actual heat con- 
sumption may be but 1 to 1.5 per cent. Compound direct-acting 
pumps running non-condensing use from 50 to 100 pounds of steam 
per i.h.p. hour. Single-cylinder flywheel pumps of the slow-speed type, 
running non-condensing, use about 50 pounds of steam per i.h.p. hour. 
Multi-cylinder flywheel pumps of the high-duty type use about 25 pounds 
per i.h.p. hour when running non-condensing, and as low as 10 pounds 
when operating condensing. High-grade direct-connected motor-driven 



598 



STEAM POWER PLANT ENGINEERING 







PUMPS 



599 



power pumps have a mechanical efficiency from line to water load, at 
normal rating, of about 80 per cent. The efficiency of geared pumps at 
normal rating varies with the character of the gearing and the degree of 
speed reduction, and may range anywhere from 40 to 70 per cent. 

The steam consumption of all direct-acting boiler pumps decreases 
with the increase in speed. This is illustrated by curve B, Fig. 383, 
plotted from the tests ofal2X7JXl2 direct-acting single-cylinder 
pump at Armour Institute of Technology, and curve A based on experi- 
ments with a 16 X 12 duplex fire pump at Massachusetts Institute of 
Technology. 

Fig. 384 gives the details of the performance of a 12 X 7 J X 12 Marsh 
boiler-feed pump at the Armour Institute of Technology. 



400 



W 

Ah' 

2 300 
u 

A 

A 

a 200 
o 

1 









1 














Effect of Speed 

on the 

Economy of Small Direct-Acting 

Steam Pumps 










A 16 x 
B 12 x 


10 x 12 

7Xx 12 


Duplex 

Simplex 








\b 




















_ 




-— ^___A__ 


• 











3 ioo 



25 50 75 100 125 150 175 200 

Number of Single Strokes Per Minute 
Fig. 383. 

The determination of the power consumption of a boiler-feed pump 
is best illustrated by the following example. 

Example: A small direct-acting duplex pump uses 150 pounds of 
steam per i.h.p. hour. Gauge pressure 150 pounds per square inch; 
feed-water temperature 64 degrees F. Required the per cent of rated 
boiler capacity necessary to operate the pump. 

The head pumped against, 150 pounds per square inch, is equivalent 
to 150 X 2.3 = 345 feet of water. 

The friction through the valves, fittings, and pipe, and the vertical 
distance between suction and feed-water inlet, are assumed to be equiva- 
lent to 20 per cent of the boiler pressure, giving a total head of 150 + 
30 = 180 pounds per square inch, or 414 feet of water. 

A boiler horse power, taking into consideration leakage losses and 
the steam used by the feed pump, will be equivalent to the evaporation 



600 



STEAM POWER PLANT ENGINEERING 



of approximately 32 pounds of water per hour from a feed temperature 

of 64 degrees F. to steam at 150 pounds gauge. 

The actual work done in pumping 32 pounds of water against a head 

of 414 feet is 

414 X 32 = 13,248 foot-pounds. 



YWU 














Curves of Performance 




























OI 

Marsh Steam "Pump 

for 

Varying Speed 

Size of Pump- 12"x 7&"x ft" 

Cap. 216 Gal. per Min. at 100 Strokes 


































6000 


-600 










































1 
















































































5000 






























A 














\ 






















// 










VU 


§ 






\ 




















/c 














a 




1 


\ 


& 


































tf 4000 


-400 
* 






<k 


\ 




























u 

CO 








\ 




























o 


-15- 


*4 
g 


H 


\ 






\ 












4& 














2 


P 

M 3000 


® 


\ 








| 






• 


^ 


















w 




300 


\ 


































a 




\ 
































fc 










\ 














































































2000 


^200 




























s 


/ 








































































S**' 
































^ 


e ^ 


^ 


















1000 


-100 











































































































































































































10 



40 50 60 

Single Strokes per Min. 

Fig. 384. 



This corresponds to 

13,248 



60 X 33,000 



= 0.0067 horse power. 



■a* 



a* 



90 



100 



The total heat of one pound of steam above 64 degrees F. is 1163 
B.t.u. The heat delivered to the pump per i.h.p. hour is 



1163 X 150 = 174,450 B.t.u. 



PUMPS 601 

The amount used by the pump for each boiler horse power, disregard- 
ing efficiency, is 

174,450 X 0.0067 = 1168 B.t.u. per hour. 

The mechanical efficiency of the average feed pump ranges from 50 
to 85 per cent, depending upon its condition and the number of strokes 
per minute. Assuming it to be 65 per cent, the heat used by the pump 
per hour to deliver 32 pounds of water into the boiler is 

1168 ^ 0.65 = 1796 B.t.u. 

A boiler horse power is equivalent to 33,479 B.t.u. per hour. There- 
fore the per cent of boiler output necessary to operate the pump is 

100 X 00-470, = 5 ' 36 P er cent * 

If the exhaust steam is used for heating the feed water, the steam con- 
sumption will be 0.73 per cent of the boiler capacity, thus: The weight 
of steam consumed per boiler-horse-power hour 

1796 1 KA a 

—no = I- 54 pounds. 

Allowing a 10 per cent loss, the heat in the exhaust available for 
heating the feed water is 

[1150 - (64 - 32)] 0.9 X 1.54 = 1550 B.t.u. 

1796 — 1550 = 246 B.t.u., or the net heat required by the pump 
per hour to deliver 32 pounds of water to the boiler. 

The per cent of boiler output necessary to operate the pump is 

100 3|S-9 = - 73 - 

Pump performances are generally given in terms of the foot-pounds 
of work done by the water piston per thousand pounds of dry steam or 
per million B.t.u. consumed by the engine, thus: 

1 D t = Foot ~Pounds of work done 
Weight of dry steam used 

n , Foot-pounds of work done w 1 ___ ___ /00 _. 

2. Duty =7^-n £ ft — 7 -, -j X 1,000,000. (235) 

lotal number 01 neat units consumed 



602 



STEAM POWER PLANT ENGINEERING 







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PUMPS 



603 



(See A.S.M.E. code for conducting duty trials of pumping engines, 
Trans. A.S.M.E., 12-530, 563.) See also, Appendix F. 

Example: A compound feed pump uses 100 pounds of steam per 
i.h.p. hour; indicated horse power, 48; capacity, 400 gallons per minute; 
temperature of water, 200 degrees F.; total head pumped against, 
175 pounds per square inch; steam pressure, 100 pounds gauge; moisture 
in the steam, 3 per cent. Required the duty on the dry steam and on 
the heat-unit basis. 

175 pounds per square inch is equivalent to 175 X 2.4 — 420 feet of 
water at 200 degrees F. 

Weight of 400 gallons of water at 200 degrees F. = 400 X 8.03 = 
3212 pounds. 

Work done per minute = 3212 X 420 = 1,349,040 foot-pounds. 

Weight of dry steam supplied per minute 



100 X 48 
60 



X 0.97 = 77.6 pounds. 



B.t.u. supplied per minute 
100 X 48 



60 



(0.97 X 879.8 + 309 - 200 + 32) = 79,552. 



Duty per thousand pounds of dry steam 
1,349,040 



77.6 



X 1000 = 17,384,150 foot-pounds. 



Duty per million B.t.u. 
1,349,040 



79,552 



X 1,000,000 = 16,958,000 foot-pounds. 



Table 96 may be used in approximating the duty, thus: 

The mechanical efficiency of the pump in the preceding problem is 



Efficiency = 



p.h.p. _ 1,349,040 



i.h.p. 33,000 X 48 



= 85 per cent. 



At the intersection of vertical column "85" and horizontal column 
" 100 " of Table 96, we find 16.82 millions. See also, Table 65. 

Tables 97 and 98 give the maximum theoretical height to which 
pumps may lift water by suction at different temperatures. In prac- 
tice these figures cannot be realized. It is customary to have the water 
gravitate to the pump for all temperatures over 120 degrees F. 



604 



STEAM POWER PLANT ENGINEERING 



TABLE 97. 

MAXIMUM HEIGHT TO WHICH PUMPS CAN RAISE WATER BY SUCTION. 
(Temperature of Water 40 degrees F.; Barometer 29.92.) 



Vacuum in 






Vacuum in 






Suction Pipe, 
Inches of 


Theoretical 
Lift. 


Probable 

Actual 

Lift. 


Suction Pipe, 
Inches of 


Theoretical 
Lift. 


Probable 

Actual 

Lift. 


Mercury. 




Mercury. 






Feet. 


Feet. 




Feet. 


Feet. 


1 


1.1 


0.9 


16 


18.0 


14.4 


2 


2.2 


1.8 


17 


19.1 


15.3 


3 


3.3 


2.7 


18 


20.2 


16.1 


4 


4.5 


3.6 


19 


21.4 


17.1 


5 


5.6 


4.5 


20 


22.5 


18.0 


6 


6.7 


5.4 


21 


23.7 


18.9 


7 


7.9 


6.3 


22 


24.8 


19.8 


8 


9.0 


7.2 


23 


25.9 


20.7 


9 


10.1 


8.1 


24 


27.0 


21.6 


10 


11.3 


9.0 


25 


28.2 


22.7 


11 


12.4 


9.9 


26 


29.3 


23.9 


12 


13.5 


10.8 


* 27 


30.4 


24.3 


13 


14.6 


11.7 


28 


31.6 


25.2 


14 


15.8 


12.6 


29 


32.7 


26.1 


15 


16.9 


13.5 


t 29.68 


33.6 









* Vacua greater than 27 inches are practically unobtainable in pumping practice except in 
connection with condensers. 

t Maximum theoretical vacuum obtainable with water at 40 degrees F. and barometer of 
29.92 inches. 



TABLE 98. 

MAXIMUM THEORETICAL HEIGHT TO WHICH A PUMP CAN LIFT WATER BY 

SUCTION AT DIFFERENT TEMPERATURES. 

(Barometer 29.92.) 



Temperature of 


Maximum 


Temperature of Feed 


Maximum 


Feed Water. 


Theoretical Lift. 


Water. 


Theoretical Lift. 


Degrees F. 


Feet. 


Degrees F. 


Feet. 


40 


33.6 


130 


29.2 


50 


33.5 


140 


27.8 


60 


33.4 


150 


25.4 


70 


33.1 


160 


23.5 


80 


32.8 


170 


20.3 


90 


32.4 


180 


16.7 


100 


31.9 


190 


12.8 


110 


31.3 


200 


7.6 


120 


30.3 


210 


1.3 



I 



i 



PUMPS 605 

310. Size of Boiler-feed Pump. 

Let D = diameter of water cylinder, inches. 

d = diameter of the steam cylinder, inches. 

L = length of stroke, inches. 

N = number of working strokes per minute. 

H = head in feet between suction and boiler water level. 

R = resistance in pounds per square inch between suction level 
and boiler water level due to valves, pipes, and fittings. 

p = boiler pressure, pounds per square inch. 

S = ratio of the water actually delivered to the piston displace- 
ment. 

W = weight of water delivered, pounds per hour. 

/ = indicated horse power of the pump at maximum capacity. 

E = mechanical efficiency of the pump, taken as the ratio of the 
water horse power at the discharge opening to the indicated 
horse power of the pump, steam end. 



Then 



W = V Hi ' if X 60 X 62 ' 5 XS = h7 D2LNS - (236) 



D= °- 77 \ms' (237) 



d=D\J 



P + R + 0AS3H 

—Ep < 238 ) 

= W(p + R + 0.433 H) 2.3 . 

33,000 X 60 X E K - } 

In average practice the piston or plunger displacement is made about 
twice the capacity found by calculation from the amount of water 
required for the engine, to allow for leakage, steam consumption of the 
auxiliaries, blowing off, and pump slip. 

For pumps with strokes of 12 inches or over, the speed of the plunger 
or piston is usually limited to 100 feet per minute as a maximum to in- 
sure smooth running. For shorter strokes a lower limit should be used. 
The maximum number of strokes ranges from 100 for strokes over 12 
inches in length, to 200 for strokes under 5 inches. Boiler-feed pumps 
should be designed to give the desired capacity at about one-half the 
maximum number of strokes or less. 

Pump slip varies from 2 to 40,per cent, depending upon the condition 
of the piston and valves and the number of strokes. An average value 
for piston and plunger pumps in first-class condition is 8 per cent when 
operating at rated capacity, but it is wise to allow a much larger figure, 
say 20 per cent, for leakage caused by wear. 



606 STEAM POWER PLANT ENGINEERING 

The area of the steam cylinder is made from 2 to 2.5 times that of 
the water end to allow for the various friction losses and the drop in 
pressure between the pump throttle and the boiler. The total head 
pumped against includes the suction lift, the friction of valves and 
fittings, the distance between the suction inlet and the boiler level and 
the boiler pressure. The excess head varies in practice from 15 to 40 
per cent of the boiler pressure; an average figure is 25 per cent. In 
allowing for the drop in steam pressure between boiler and pump a liberal 
figure is 25 per cent. 

The application of formulas (236) to (239), including the practical 
considerations stated above, is best illustrated by a specific example. 

Example: Determine the size of direct-acting single-cylinder feed 
pump necessary to supply water to 1000 horse power of boilers. Gauge 
pressure 100 pounds per square inch; feed-water temperature 150 
degrees F. 

One horse power is equivalent to the evaporation of 34.5 pounds of 
water from and at 212 degrees F.; but the pump is usually designed to 
supply about twice the capacity. 

Thus W = 62,400 (under the given conditions). 
S = 0.8 (by assumption). 
LN = 1200 (on the basis of 100 feet per minute). 

Substitute these values in (237) : 



D = 0.77 v " ? = 6.2 inches, — call it 6 inches, 



1200 X 0.8 

since the assumptions have been very liberal. 
Assume (0.433 H + R) = 0.25 p and E = 0.65. 

Substitute these values in (238) : 



-«v/„ J 



100 + 25 



.65 X 100 
= 8.35, — call it 8.5 inches. 

Allowing 100 strokes per minute the length of the stroke must be 
L = 1200 -f- 100 = 12 inches. 

The dimensions of the pump are 8 J X 6 X 12. 

The indicated horse power at maximum load may be obtained by 
substituting the proper values in (239), thus: 

= 62,400 (100 + 25) 2.3 

33,000 X 60 X 0.65 
= 13.9 i.h.p. 



PUMPS 



607 



311. Steam-pump Governors. — Fig. 385 shows a section through a 
Fisher pump governor, illustrating a device for maintaining a practically 
constant pressure in the discharge pipe irrespective of the quantity of 
water flowing. It embodies a pressure-reducing valve in the steam 
supply pipe of the pump, actuated by the slight E 
variations in water pressure. When the demand 
for water increases, the pressure in the discharge 
pipe tends to decrease, and this drop in pressure 
(transmitted to the pump governor by suitable 
piping) causes more steam to be admitted, which 
increases the speed of the pump. The governor 
is connected to the steam inlet of the pump at B 
and the steam enters at A . Double-balanced valve 
C regulates the supply of steam to the cylinder by 
the amount it is raised from the seat. The valve 
is held open by spring G, the compression of which 
may be regulated by hand wheel K. The water 
pressure from the discharge pipe acts on piston F 
and tends to overcome the resistance of the spring. 
The difference in pressure between the water and 
the spring determines the position of valve C. 

Piston rod H is pinned to sleeve 7 and valve 
stem L screwed into this sleeve by means of hand 
wheel K. Hence, during ordinary operation, the 
piston, piston-rod sleeve, valve stem, and valve 
act as a single unit. By turning the hand wheel 
K, valve stem L will screw into sleeve I and the 
tension on the spring will be increased. Hand 

wheel J serves as a lock nut and prevents K from turning during 
normal operation. 

312. Feed-water Regulators. — The water level in the boiler should 
be kept as nearly constant as possible, and this necessitates considerable 
attention on the part of the fireman, especially with fluctuating loads. 
There are a number of devices on the market which are designed to 
automatically maintain a constant level, and in many small plants 
where the duties of the fireman are numerous such devices in connec- 
tion with high and low water alarms are of considerable assistance. 
Their action, however, is not always positive on account of wear or 
sticking of parts, and engineers as a rule prefer to rely upon hand regu- 
lation. 

Fig. 386 shows a section through a Kitts feed-water regulator, con- 
sisting of two parts, the chamber F and the regulating valve V. The 




Fig. 385. Fisher Pump 
Governor. 



608 



STEAM POWER PLANT ENGINEERING 



float chamber is connected to the boiler or water column at and E, 
and the regulating valve to the feed main at R and to the boiler feed 
pipe at W. When the water in the boiler falls below the mean level, the 
weight B overcomes the counterweight G and closes needle valve L by 
means of compound levers. At the same time an extension on valve L 
lifts spring A and opens exhaust valve D. This removes the steam 
pressure from the top of diaphragm C, in the regulating valve, through 
the agency of pipe K. The pressure from the pump raises the disk T 
and water flows into the boiler until the water rises to the mean level. 





Fig. 386. Kitts Feed-water Regulator. 



Fig. 387, Rowe Feed-water 
Regulator. 



When weight B becomes submerged its weight is overcome by counter- 
weight G, valve L is opened and exhaust valve D is closed. This 
admits steam pressure to the diaphragm C and forces disk T to its seat, 
cutting off the supply of water to the boiler. 

The Rowe feed-water regulator, Fig. 387, depends for its operation 
on a familiar float-controlled valve mechanism. The vessel A is con- 
nected to the boiler above and below the water line, and the float C, 
following the water level up and down, actuates a balanced valve in 
accordance with the boiler-feed requirements. When this apparatus 
is used to regulate the feed of a single boiler the opening G in the valve 
chamber is connected to the steam space of the boiler and the outlet H 
is carried to the steam inlet of the feed-water pump. When the water 



PUMPS 



609 



level is normal the float closes the valve L and thereby cuts off the 
supply of steam to the pump cylinders. Communication between 
chambers A and R is prevented by means of a diaphragm M . When 
the water level falls below normal the float pulls the valve down, open- 
ing the way for steam to pass from the inlet G to the outlet H and 
thence to the pump. When the regulator is used to control a battery 
of boilers the pump discharge delivers into the inlet G and the water 
passes through H to the boiler-feed main. Should the water level fall 



b~ 


w 


=1 

P 

r\ 

E 




Fig. 388. " S-C " Feed Water Regulator. Fig. 389. A Typical Geared Triplex Pump. 

beyond a predetermined limit by reason of any accidental discon- 
tinuance of the water supply which the apparatus cannot correct, the 
float would open the valve F of the alarm whistle mounted on the top 
of the main vessel. ^ 

Fig. 388 gives the general details of the "S-C" feed water regulator 
which differs from the types just described in the manner of actuating 
the water regulating valve. A small copper vessel, A, partly filled 
with water, is in communication with diaphragm / through the medium 
of tube F. The water in vessel A is independent of the boiler supply. 
A small copper U-tube, U, projects into chamber A, as indicated. When 
the water in the boiler is at its highest level the U-tube is filled with water 



610 



STEAM POWER PLANT ENGINEERING 



and the pump regulator valve V is not feeding. As the level of the 
water in the boiler drops the water recedes from the outer surface of 
the U-tube, and the upper branch of the tube is surrounded with steam. 













































1 






































































































Knowles High Speed Electric Pump 

Direct connected to 

M P 6-100 H.P. -280-220 V.Form L 

Load and Efficiency 
































































































































































































































I 






1 


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Noz 


de 


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orizoi 


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-2: 


o 




















Gall 


on 


sP 


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Minu 


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/ 




































































-£ 






















-2L 
















































r 1 




































































/ 


































































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„ 












































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-260- 








































^ 


f/ 
































































&/ 






































































A 
































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10- 












































V 
































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ri 
































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Pl 














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100 






200 






300 






400 






500 




600 




























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Gauge Pressure at Valve (Lb.) 
Fig. 390. 



The steam causes the water in the vessel A to boil, and the pressure 
generated is transmitted through pipe F to diaphragm I, thereby open- 
ing controlling valve K. Wheel J permits of hand control. Regulators 



PUMPS 



611 



of this type installed in the power plant of the Armour Institute of Tech- 
nology are giving excellent service. 

313. Power Pumps. - — Piston pumps, geared, belted, or direct con- 
nected to electric motors, gas engines, and water motors, are used 



100 



75 



I" 



O 50 











Motor _ 
' Pump 






md Gearing 


















Set . 


















Head Cons 


ant, 350 Ft. 



















100 






50 



300 400 600 800 1000 

Discharge, Gallons Per Minute 
Head Constant. Speed Variable. 



1200 



1400 









Motor 










IPUEttPji 


ad Gearing 


















Set 
















^ 






10 x 


2 Triplex P 
5 H.P. Mote 


imp 
r 










Mc 
C 

Capacit 


tor B.P.M. 
earing 10 to 
7 1250 Gal. p 


150 

;r Min. 



25 50 75 100 125 

Total Head, Lb. PerSq. In. Gauge 
Speed Constant. Head Variable. 



150 



175 



Fig. 391. Performance of a 65-horse-power, Motor-driven Triplex Pump. Geared Type. 

chiefly where steam power is not available. Their general utility is 
evidenced by the rapidly increasing number installed in situations 
formerly occupied by the direct-acting steam pump. The efficiency of 
this type of pump depends in a large measure upon the character of 
the driving motor and the efficiency of the transmitting mechanism. 
High-speed power pumps direct connected to electric motors give 



612 



STEAM POWER PLANT ENGINEERING 



efficiencies from line to water horse power as high as 83 per cent, while 
the low-speed geared type seldom exceed 70 per cent. The curves in 
Fig. 390 give the performance of a direct-connected triplex pump, 
and those in Fig. 391 the performance of a triplex pump geared to an 
electric motor. Both of these performances are exceptionally good 
and are considerably above the average. 

For a General Treatise on the Design and Operation of Pumping Machinery con- 
sult "Pumping Machinery," by A. M. Greene; John Wiley & Sons, 1911. 

314. Injectors. — As a boiler feeder the injector is an efficient and 
convenient device, cheap and compact, with no moving parts, delivers 
hot water to the boiler without preheating, and has no exhaust steam 
to be disposed of. Its adoption in locomotives is practically universal, 
but in stationary practice it is limited to small boilers or single boilers 
or as a reserve feeder in connection with pumps. The objections to 



Steam Supply 




Boilel 



Overflow 

Fig. 392. Elementary Steam Injector. 

an injector are its inability to handle hot water, the difficulty of main- 
taining a continuous flow under extreme variation of load, and the un- 
certainty of operation under certain conditions. Fig. 392 illustrates 
the simplest form of single-tube injector. Boiler steam is admitted at 
A and, flowing through nozzle and combining tube to the atmosphere 
through G, partially exhausts the air from pipe B, thereby causing the 
water to rise until it comes in contact with the steam. The steam 
emerging from nozzle C at high velocity condenses on meeting the 
water and imparts considerable momentum to it. The energy in the 
rapidly moving mass is sufficient to carry it across opening 0, lift check 
H from its seat and force it into the boiler. The steam then ceases 
to escape at G. 

315. Positive Injectors. — Fig. 393 shows a section through a Han- 
cock injector, illustrating the principles of the double-tube positive 
type. Its operation is as follows: Overflow valves D and F are opened 
and steam is admitted, which at first passes freely through the over- 
flow to the atmosphere and in so doing exhausts the air from the suction 
pipe. This causes the feed water to rise until it meets the jet of steam 



■ I 



PUMPS 



613 



and the two are forced through the overflow. As soon as water appears 
at the overflow, valve D is closed, valve C partially opened, and valve 
F closed. This admits steam through the forcing jet W and, the over- 
flow valves being closed, the water is fed into the boiler. In case the 
action is interrupted for any reason it is necessary to restart it by hand. 
The chief advantage of the double-tube positive type lies in its 
ability to lift water to a greater height and to handle hotter water than 
the single-tube. Its range in pressure is also greater, that is, it will 
start with a lower steam pressure and discharge against a higher back 
pressure. Double-tube injectors are used almost exclusively in loco- 
motive work. 




Overflow 

Fig. 393. Hancock Double-tube 
Injector. 




Fig. 394. 



Penberthy Automatic 
Injector. 



316. Automatic Injectors. — Fig. 394 shows a section through the 
Penberthy injector. Its operation is as follows: Steam enters at the 
top connection and blows through suction tube c into the combining 
tube d and into chamber g, from which it passes through overflow valve 
n to the overflow m. When water is drawn in from the suction intake 
and begins to discharge at the overflow, the resulting condensation of 
the steam creates a partial vacuum above the movable ring h and the 
latter is forced against the end of tube c, cutting off the direct flow 
of water to the overflow. The water then passes into the boiler. Spill 
holes i, i, i are for the purpose of relieving the excess of water until 



614 STEAM POWER PLANT ENGINEERING 

communication with the boiler has been established. The action of 
opening and closing the overflow is entirely automatic. Where the 
conditions are not too extreme the automatic injector is to be preferred 
for stationary work because of its restarting features. It is also used on 
traction, logging, and road engines, where its certainty of action and 
special adaptability render it invaluable for the rough work to which 
such machines are subjected. 

Injectors, Theory of: Trans. A.S.M.E., 10-339; Sibley Jour., Dec, 1897, p. 101; 
Power, May, 1901, p. 23; Thermodynamics of the Steam Engine, Peabody, Chap. 
IX; Theory of the Steam Injector, Kneass. 

Injectors, General Description: Engr. U. S., Oct. 1, 1907, Nov. 15, 1907, July 15, 
1904, p. 501, Feb. 2, 1903, p. 151; Power, Aug., 1906, p. 478; Engr., Lond., March 
10, 1905, p. 244; Engineering, Aug. 30, 1895, p. 281. 

317. Performance of Injectors. — The performance of an injector may 
be very closely determined from the equation 

_ xr + q — t + 32 (Kneass, "Theory of the (oac\\ 

t -U Injector," p. 83), ( ] 

in which 

w = pounds of water delivered per pound of steam supplied, 
x = quality of the steam supplied, 
r = heat of vaporization, 
q = heat of the liquid, 
t = temperature of the discharge water, 
t = temperature of the suction water. 

Figs. 395, 396, and 397 give the performance of a Desmond auto- 
matic injector as tested at the Armour Institute of Technology. The 
results check very closely with those calculated from above equation. 
Referring to Fig. 395 it will be seen that the weight of water delivered 
per pound of steam decreases as the initial pressure is increased, all 
other factors remaining the same. From Fig. 396 it will be noted that 
the weight of water delivered per pound of steam decreases as the tem- 
perature of suction supply is increased up to a point where the injector 
" breaks" or becomes inoperative. This critical temperature varies 
with the different types of injectors, being highest for the double-tube 
type, but seldom exceeds 160 degrees F. Fig. 397 shows that the weight 
of water delivered per pound of steam is practically constant for all 
discharge pressures within the limits of the apparatus. 

Table 99 gives the range of working steam pressures for standard 
"Metropolitan" injectors with varying suction heads and temperatures, 
and, though strictly applicable to this particular type only, is charac- 
teristic of all makes. 



PUMPS 



615 



«3 -2 



21 



20 



19 



IS 



17 



a 



13 







































Constant Discharge Pressure 

20 Lb. Per Sq.In. 

Constant Suction Temp. 

55 Deg.Fah. 






























































ON 
































v () 







































































C5 



75 80 85 

Initial Gauge Pressure.Lb.Per Sq.In. 



90 



95 



Fig. 395. Performance of an Automatic Injector with Varying Initial Pressure. 



a* 
*° _ 

>! 

«<3 



2U 
19 
IS 
17 
16 
,15 
14 








































Constant Initial Pressure 

70 Lb. Per Sq. In. 

Constant Discharge Pressure 

70 Lb-Per Sq. In. 








C 


>^ 






















c 


D^^"- 










































C 

































65 



75 85 95 

Temperature of Suction, Deg. Fan. 



105 



115 



Fig. 396. Performance of an Automatic Injector with Varying Suction Temperature. 



2 a 
£ a 

5> 02 



la 



20 



19 



18 



17 



16 





Constant Initial Pressure,70 Lb. Per Sq. In. 
Constant Suction Temperature, 56 Deg. Fan. 












c 


)— 








( 


),- 






















k 


) 










1 


1 


( 


J 

































30 .40 50 60 

Discharge Pressure, Lb. Per Sq. In. Gauge 



70 



80 



Fig. 397. Performance of an Automatic Injector with Varying Discharge Pressure. 



616 



STEAM POWER PLANT ENGINEERING 



In selecting an injector the following information is desirable for 
best results: 

1. The lowest and highest steam pressure carried. 

2. The temperature of the water supply. 

3. The source of water supply, whether the injector is used as a 
lifter or non-lifter. 

4. The general service, such as character of the water used, whether 
the injector is subject to severe jars, etc. 

Injectors, Tests of: Eng. News, March 17, 1898, July 16, 1896, p. 39; Locomotive 
Engineering, May, 1900, p. 204; Power, Oct., 1904, p. 602; Railroad Gazette, Dec. 
11, 1896; Thermodynamics of the Steam Engine, Peabody, Chapter IX; Theory of 
the Injector, Kneass. 

TABLE 99. 

RANGE IN WORKING PRESSURES. 
Standard " Metropolitan " Steam Injectors. 





Automatic. 


Suction 
Temperature, 


Suction Head, Feet. 


Degrees F. 


2 


8 


14 


20 


Under 
Pressure. 


Under 60 


25 to 150 


30 to 130 


42 to 110 


55 to 85 


20 to 160 


100 


26 to 120 


33 to 100 


55 to 80 




25 to 125 








120 










26 to 85 














140 































Double Tube. 






Suction 
Temperature, 


Suction Head, Feet. 


Degrees F. 


2 


8 


14 


20 


Under 
Pressure. 


Under 60 


14 to 250 


23 to 220 


27 to 175 


42 to 135 


14 to 250 


100 


15 to 210 


26 to 160 


37 to 120 


46 to 70 


15 to 210 


120 


20 to 185 


30 to 120 


42 to 75 




20 to 185 








140 


20 to 120 


35 to 70 






20 to 120 











PUMPS 617 

318. Injector vs. Steam Pump as a Boiler Feeder. — From a purely 
thermodynamic standpoint the efficiency of an injector is nearly per- 
fect, since the heat drawn from the boiler is returned to the boiler again, 
less a slight radiation loss. As a pump, however, the injector is very 
inefficient and requires more fuel for its operation than very wasteful 
feed pumps. This is best illustrated by an example: An injector of 
modern construction will deliver say 15 pounds of water to the boiler 
per pound of steam supplied, with delivery temperature of 150 degrees F. 
This corresponds to a heat consumption of 71.3 B.t.u. per pound of 
water delivered, thus: 

With initial pressure of 115 pounds absolute, 

H = 1188.8. 

Heat in the water delivered to the boiler, 

150 - 32 = 118 B.t.u. above 32 degrees F. 

Heat of 1 pound of steam above a feed temperature of 150 degrees F., 
1188.8 - 118 = 1070.8 B.t.u. 

Heat required to deliver 1 pound of water to the boiler, 

1070.8 71 ', 
— — — =71.3 B.t.u. 
15 

A simple direct-acting duplex pump consumes say 200 pounds steam 
per i.h.p. hour. Assume the extreme case where the exhaust steam 
will not be used for heating the feed water and the latter is fed into 
the boiler at 60 degrees F. 

The heat supplied to the pump per i.h.p. hour, 

200 51188.8 - (60 - 32) J = 232,160 B.t.u. 

Assuming the low mechanical efficiency of 50 per cent, the heat 
required to develop one horse power at the water end will be 
232,160 -h 0.50 = 464,320 B.t.u. per hour. 

Since the steam pressure is 100 pounds gauge, the equivalent head 
of water at 60 degrees F. is 

2.3 X 100 = 230 feet. 

Assume the friction in the feed pipe, the resistance of valves, etc., to 
be 30 per cent of the boiler pressure; the total head pumped against 
will be 

230 + 69 = 299, say 300 feet, 

1 horse-power hour = 1,980,000 foot-pounds per hour, 
1,980,000 



300 



6600 pounds; 



618 STEAM POWER PLANT ENGINEERING 

that is, 1 horse power at the pump will deliver 6600 pounds of water per 
hour to the boiler against a head of 300 feet. 

The heat consumption per pound of water delivered, 

464,320 



6600 



= 70.3 B.t.u. 



If the feed water is heated to say 210 degrees F. by the exhaust steam 
from the pump, the heat consumption will be 63.7 B.t.u. as against 70.3 
without the heater. 

Thus even in this extreme case of poor steam-pump performance 
the heat consumption lies in favor of the pump. With the better 
grades of pumps this disparity is considerably greater, and decidedly 
so if the exhaust steam is used to preheat the feed water. For inter- 
mittent operation the condensation losses in the pump may more than 
offset this gain. Other conditions, however, such as compactness, 
low first cost, and ease of operation are oftentimes considerations and 
the heat consumption is of minor importance. 

319. Air Pumps. — Condenser air pumps may be divided into two 
classes : 

1. Wet-air pumps and 

2. Dry-air pumps. 

The former handle both air and water and the latter air alone. Or- 
dinary jet-condenser wet-air pumps handle simultaneously the circu- 
lating water, condensed steam, and entrained air, and are, in fact, a 
combination of circulating and vacuum pump. Surface-condenser 
wet-air pumps are the same in principle and design, but are smaller in 
size for a given main engine output, as they handle the condensed steam 
and air only. 

Wet-air pumps may be driven by the main engine or independently 
and may be direct acting, Fig. 296, or flywheel driven, Fig. 312. The 
flywheel type may be steam, electric, or belt driven. Dry-air pumps 
are virtually air compressors, as their function is to compress air from 
the pressure existing in the condenser to that of the atmosphere. They 
are generally of the flywheel type. 

320. Dean Air Pump. — Fig. 398 shows a section of the air cylinder 
of a Dean twin-cylinder wet-air pump as applied to a standard low- 
vacuum jet condenser. There are three sets of valves, the suction or 
foot valves A, A, the lifting or bucket valves B, B, and the head or 
discharge valves C, C. On the upward stroke of the piston or bucket 
a partial vacuum is formed in the chamber between the bucket and the 
lower head, causing the water and air in the bottom of the barrel to 



PUMPS 



619 




lift the foot valves A, A from their seats and flow into the cylinder. 
On the downward stroke the foot valves A, A close and water and air 
are entrapped in chamber R between 
the lower head and the bucket. 
As the bucket descends, the pres- 
sure of air in the cylinder lifts the 
bucket valves B, B from their seats 
and permits the air and water to 
escape to the upper portion $ of 
the cylinder between the head plate 
and the bucket. On the next up- 
ward stroke the water and air are 
forced through the discharge valves 
C, C into the hot well. This dis- 
charge of water and air from the top 
compartment is simultaneous with 
influx of water and air in the lower 
chamber. 

See paragraph 242 for other types 
of wet-air pumps in connection with 
jet condensers. 

321. Size of Wet-air Pumps; Jet Condensers. — In proportioning such 
pumps the quantity of cooling water and condensed steam to be 
taken care of is readily determined, but the percentage of air mingled 
with it must be estimated. Surface water under atmospheric pressure 
ordinarily contains from 2 to 5 per cent of air by volume. To provide 
for possible leakage a very liberal factor is usually allowed, an average 
figure being about 10 per cent. 

Let Q = total volume of air and water in cubic feet per hour to 
be handled by the pump; 

V = volume of cooling water in cubic feet per hour; 

v = volume of condensed steam in cubic feet per hour; 

v a = volume of air at pressure p a and temperature t a ; 

t a — temperature of the air entering the condenser, degrees F.; 

to, = temperature of the discharge water, degrees F. ; 

t = initial temperature of the cooling water, degrees F. ; 
p a = atmospheric pressure, pounds per square inch; 
p c = total pressure in the condenser, pounds per square inch; 
p v = pressure of aqueous vapor at temperature t 2 ; 



Fig. 398. Dean Air Pump. 



then (V + v) 
per hour. 



volume of water to be pumped from the condenser 



620 



STEAM POWER PLANT ENGINEERING 



The air entering the condenser will be increased in volume on account 
of the reduction in pressure and the increase in temperature. If v a is 
the original volume under pressure p a and temperature t a the final 
volume on entering the condenser is (see equations 163 and 164) 



Final volume 



Va' 



„ t2 + 460 



Vc — Vv t a + 460 
and the total volume to be exhausted per hour by the pump is 



Q = V + v + v 



Pc 



X 



h + 460 



Vc — Vv t a + 460 



(241) 



(242) 



Exhaust 



Condenser 




Fig. 399. 



Main Pump Dry Air Pump 

Hewes and Phillips Jet Condenser and Air Pumps. 



Under average conditions of reciprocating-engine practice the hot- 
well temperature is about 110 degrees F. and the absolute back pressure 
4 inches of mercury. Assuming 70 degrees F. as the initial temperature 
of the circulating water and allowing 10 per cent as the air entrainment, 

Va = 29.92 t = 70 v = 0.04 7 (see equation 167) 

Vc = 4 t 2 = 110 v a = 0.1 V. 

Vv = 2.59 ta = U = 70 

Substitute these values in (242) 

29.92 _ 110 + 460 



Q= V + 0.04 7 + 0.1 V 
= 3.3 7. 



4.0 - 2.59 X 70 + 460 



PUMPS 



621 



Average practice gives 3 V as the pump displacement per hour for a 
single-acting pump and 3.5 7 for a double-acting pump, the cylinders 
being ordinarily proportioned on a piston velocity of 50 feet per minute 
at rated capacity. 

Table 100 gives the approximate sizes of air pumps for condensers 
as manufactured by prominent makers. 

The combined air and circulating pump is not adapted for high 
vacuum work on account of the enormous increase in air volume at 
very low pressures. With cold injection water and a good air-tight 
condensing system vacua as high as 2 inches absolute are possible 
with the standard type of jet condenser air pumps but practice recom- 
mends the use of separate air and circulating pumps under these con- 
ditions. (See paragraph 326.) 

TABLE 100. 

APPROXIMATE SIZES OF PISTON AIR PUMPS FOR STANDARD LOW VACUUM 

CONDENSERS. 



Pounds of 
Steam Con- 
densed per 
Hour. 


Jet Condenser. 


Surface Condensers. 


Duplex 
Pump. 


Horizontal 

Double Acting 

Pump. 


Vertical 

2-Cylinder 

Single 

Acting. 


Horizontal. 


Vertical 
2-Cylinder. 


500 to 1,000 


4f X 5 
5f X 6 
6|X6 
7*X 6 

7 X 10 

8 X 10 
8£X 10 

9 X 10 
10 X 10 
10* X 10 

11 X 10 

12 X 10 
12 X 15 
15 X 15 
17 X 15 
19 X 15 


6X 7 
8X 7 
8X 12 
9X 9 
9X 10 

11 X 12 

12 X 14 
14 X 14 

14 X 16 

15 X 16 

15 X 18 

16 X 18 
18 X 18 
20 X 24 
24 X 24 
26 X 24 


5X 4 

6X 4 

7X 5 

9X 6 

10 X 8 

11X 9 

12 X 8 

12 X 10 

14 X 10 

15 X 10 

15 X 12 

16 X 10 

17 X 12 
20 X 12 
22 X 15 
24 X 18 


3*X 4 
4X4 
4X6 
5 X 7 
5X8 
6X8 
7X9 
7 X 10 

7 X 12 

8 X 10 

8 X 12 

9 X 12 
10 X 12 
12 X 14 
14 X 16 
16 X 24 




1,000 to 1,500 




1,500 to 2,000 




2,000 to 2,500 




2,500 to 3,000 




3,500 to 4,000 




4,000 to 4,500 




4,500 to 5,000 




5,000 to 6,000 




6,000 to 7,000 

7,000 to 8,000 

8,000 to 9,000 

9,000 to 10,000 

10,000 to 15,000 

15,000 to 20,000 

20,000 to 25,000 


8X 4 
8X 6 
9X 6 
10 X 8 
11X 8 
12 X 8 
14 X 10 



Wet-air pumps are usually independently driven, making it possible 
to vary the speed of the pump irrespective of the engine speed and to 
create a vacuum before starting the engine. Occasionally, however, 
when the load is constant, as in pumping-engine practice, the pump 
may be driven by the main engine. 

Centrifugal wet-air pumps are much in evidence in modern stations and 
offer many advantages over the piston type. See paragraphs 330 to 334. 



622 



STEAM POWER PLANT ENGINEERING 



322. Edwards Air Pump. — Fig. 400 shows a section through the 
air cylinder of an Edwards air pump. This device belongs to the sur- 
face-condenser " wet-air pump" class, as both the water of condensa- 
tion and the entrained air are exhausted simultaneously by the same 

piston. Unlike the standard 
type of wet-air pumps, foot 
valves and bucket valves are 
entirely dispensed with. The 
condensed steam flows contin- 
uously by gravity from the con- 
denser into the base of the 
pump through passage A and 
annular space B. As the piston 
C descends it forces the water 
from the lower part of the cas- 
ing F into the cylinder proper 
through the ports P, P. On the 
upward stroke the ports in the 
piston are closed and the air 
and water discharged through 
head valves D and exhaust 
port E to /the hot well. The 
seats of valves D are con- 




Fig. 400. Edwards Air Pump. 



structed with a rib between each valve and a lip around the outer edge, 
so that each valve is water-sealed independently of the others. In 
earlier air pumps of this general type the clearance between the bucket 
and head valve seat is necessarily large, due to the space occupied by 
the bucket valves and the ribs on the under side of the valve seating. 
This clearance space reduces the capacity of the pump, since the air 
above the bucket must be compressed above atmospheric pressure be- 
fore it can be discharged, and on the return stroke will expand and 
occupy a space which should be available for a fresh supply of air from 
the condenser. In the Edwards air pump the clearance space is reduced 
to a minimum, since there are no bucket valves to limit it. The absence 
of suction or foot valves still further increases the capacity of the pump 
for similar reasons. These pumps are arranged either single, double, or 
triplex; steam, electric, or belt driven; slow or high speed. They are 
ordinarily used in connection with surface condensers. 
Centrifugal Wet Vacuum Pump: Power & Engr., Jan. 4, 1910. 

323. Mullan Valveless Air Pump. — Fig 401 shows a section through 
the "Mullan valveless air pump" as used in connection with the C. H. 



PUMPS 



623 



Wheeler Company's " high-vacuum " condensing outfit. The pump is 
double acting and devoid of suction valves. The cylinder has a central 
port which is uncovered by the piston at each end of the stroke and 
covered at all other positions. Discharge valves of the Gutermuth 
spiral-spring type are located in both heads of the cylinder. As the 




Fig. 401. Mullan Air Pump. 

piston moves from one end of its stroke to the other it forms a vacuum 
behind it and forces out the gases and water ahead of it; when it reaches 
the end of the stroke the central inlet port is uncovered and the vacuum 
behind the piston draws in the condensation and gases from the con- 
denser. This operation is repeated on the return stroke. 

The makers claim that the pump will operate, under shop-test con- 
ditions, within one half inch of the barometer, enabling them to guaran- 
tee a vacuum within two inches of absolute under full-load conditions 
of steam turbine operation. 

JH 

Discharge - 




Fig. 402. High Vacuum " Rotrex " Pump. 

324. C. H. Wheeler " Rotrex " Pump. — Fig. 402 shows a partial axial 
and an end section through a C. H. Wheeler & Co.'s high- vacuum 
" Rotrex" pump. This pump is of the wet-vacuum type and handles 



624 



STEAM POWER PLANT ENGINEERING 



both air and water of condensation but it is also adapted for dry-air 
purposes. The apparatus consists of a cylindrical casing and a rotor 
mounted eccentrically on the shaft. This shaft is carried in outboard 
ring oil bearings which are entirely independent of the stuffing boxes. 
The division between the suction and discharge space in the pump 
cylinder is maintained by a radius cam carried on a shaft independent 
of the stuffing boxes. This cam is operated from the rotor shaft by a 
lever and crank on the outside of the casing. The clearance spaces are 
water sealed. The discharge valves are of the Gutermuth type. Pump 
speed 200 to 300 r.p.m. The manufacturers guarantee that on dead- 
end test a vacuum may be obtained within one half inch of the barom- 
eter, and within one inch of the barometer under operating conditions. 



II 


























i 1 












A- 


*<"- 










■X- 


/ 


--A 












3 










/ 


/ 
















s 1 

o 










/ 
























\ 


/ 


















o 

XL 

o 

C 

M 


B- 






__\ 


t- 






-P 




















\ 


















U 

< 








J 


\ 












































3 

0) 












\2 
V* 
























o. 




1 1 
















3 
O 

S 3 








a 




\l 

lis* 




















a 




&S 






















3 

















































Fig. 403. 



1 2 3 4 5 6 7 8 9 10 11 12 13 
Cubic Feet of Air Removed per Second 
Comparative Tests — Reciprocating Air Pump vs. Leblanc Air Pump. 



325. Leblanc Air Pump. — The Leblanc air pump, described in para- 
graph 253, is finding much favor with engineers for high-vacuum 
service, and is supplanting the reciprocating type of pump to a con- 
siderable extent. The absence of valves, high rotative speed and the 
elimination of clearance spaces, enable large volumes of air to be handled 
at very low absolute back pressures. Since there are no reversals in 
operation no damage results from water being drawn into the air pump 
as when the hot-well pump fails to operate. For low vacua the recip- 






PUMPS 



625 



rocating air pump is more effective, volumetrically, than the Leblanc 
pump, but for high vacua the latter gives the best results. This is 
shown by the curves in Fig. 403 which, though strictly applicable only 
to the conditions under which the tests were made, represent the gen- 
eral characteristics of the two types of pumps. (Jour. Wes. Soc. of 

1500 



1400 
,1300 
1200 
1100 



^1000 

o 
•3 

3 900 
o 



* 700 
I 600 

O 

J5 500 

3 

400 
300 

200 
100 



Fig. 404. 











I 






























1 ® 
1 3 


































i 

3 


1 = 

| 3 
























tj 




■n 

\ 

4 

o 


> 

1 J3 I 
• O 
























3 
"5 
4> 


h 


1 °° 1 






















s 
a 


/ 


/ ° 




£ 




] 










£ 




4> 

in 


/ a 

1 


/ " 
/ W 
/ * 


s 

3 
3 


3 

BD 

to 

4) 

fc. 














$1 




8 1 


i 


/ « 




a 




1 


<s> 














/ 




**/ 


3 

2 




fl 


s 

2h 






/ 




< 


V 


/ 


! s 


/ 


4) 

S 




'if/ 


1 




/ 










/ 


i 
i I 


/ 


3 

1 
02 




2 


3 
















i / 
i/ 








J 


01 

a 
S 

4) 
















i 






/ 


' 


y 

4) 
00 
















i 




























i 

i 















40 50 60 70 80 90 100 110 120 130 140 150 

Temperature of Air Pump Suction, Degrees Fahrenheit 
Cubic Feet of Saturated Air Containing One Pound of Dry Air for Various 
Vacua and Air Temperatures. 



Engrs., May, 1912, p. 430.) Referring to the curves, the reciprocating 
pump is superior to the Leblanc for vacua below the line BB and for 
vacua above BB the latter is the more effective. For vacua above line 
AA the Leblanc pump is in a class by itself. The power requirement 
of the Leblanc pump is considerably more than that of the reciprocat- 
ing pump but with high vacua its volumetric capacity is so much greater 



626 



STEAM POWER PLANT ENGINEERING 



than the latter that the decreased air tension and corresponding in- 
crease in vacuum may more than offset power requirement. Table 
101 gives the results of a Leblanc condenser test and gives some idea of 
the power requirements. 

TABLE 101. 

TEST OF LEBLANC CONDENSER AND PUMP. 



Steam condensed, pounds per hour 

Approximate rating, per cent 

Ratio water to steam 

Injection temperature, degrees F 

Temperature difference between vacuum 

and discharge, degrees F 

Vacuum, inches of mercury 

Per cent of vacuum realized 

Power to operate all pumps, e.h.p ....... 



13,100 


21,100 


28,750 


37,000 


44 


70 


95 


123 


82.5 


51.4 


45.8 


33 


65 


70 


71 


70 


6 


7 


3.5 


2 


28.76 


28.09 


27.96 


27.59 


99.1 


98:6 


99.16 


99.15 


52 


53 


56 


55 



44,400 
148 

24.9 

70 

4 

26.81 

98.43 

51 



326. Size of Wet-air Pump for Surface Condenser. — Since the wet-air 
pump for surface condenser handles only the condensed steam and air, 
its theoretical capacity, neglecting clearance, may be determined by 
eliminating V from equation (242) which then becomes 

h + 460 



Q = V + V a 



Pa 



X 



Vc — Vv t a + 460 



(243) 



The volume of air entering the condenser varies so much with the 
character of the power-plant equipment and the conditions of opera- 
tion that any assumed average value of v may lead to serious error. 

Average steam turbine practice gives 

Q = 20 v for 26-inch vacuum, 
Q = 30 v for 27-inch vacuum, 
Q = 40 v for 28-inch vacuum. 

Average reciprocating engine practice gives 

Q = 85 per cent of above for same operating conditions. 

The air-pump displacement necessary to exhaust a given weight of 
air for different vacua and air-pump suction temperatures is shown in 
Fig. 404. The curves are based upon equation (241) and give the 
volume of saturated air containing one pound of dry air at various 
vacua and air-pump suction temperatures. The great reduction in 
volume effected by cooling the air-pump suction is also clearly shown. 
The marked superiority of the counter-current condenser over the 
parallel-flow condenser for high vacua is chiefly due to the reduction in 
temperature of the non-condensable vapors. 



PUMPS 



627 



327. Alberger Rotative Dry-air Pump. — Fig. 405 shows a section 
through the air cylinder of an Alberger rotative dry-air pump, illus- 
trating a type of pump in which the admission valve is mechanically 
operated. This pump is designed to operate with dry air only, all 
condensation being removed before the air enters the cylinder. This 
permits of the use of a small clearance space and makes it possible to 
run at a higher speed of rotation than can be secured with a type of 
pump in" which water is used to seal the valves. Referring to Fig. 405, 
air is being taken into the right-hand end of the cylinder through inlet 
A and forced from the left-hand end through exhaust opening B. 
Rotary valve mechanically opens to admission and mechanically 




Fig. 405. Alberger Rotative Dry-air Pump. 

closes the discharge. The discharge opening depends on the spring- 
regulated valve C at the top of the cylinder. Heads are water jacketed. 
Ports and passages are made large to reduce the friction of the air 
entering the pump, and, to obviate the bad effects of clearance, an 
equalizing passage is provided in valve 0. The action of the passage is 
shown in the section to the right. When the piston reaches the end of 
the stroke the clearance space is filled with air at atmospheric pressure. 
If this pressure were not relieved the piston would travel a considerable 
distance before drawing in air from the condenser. By means of the equal- 
izing passage the clearance space is connected to the opposite end of the 
cylinder and the vacuum there reduces the pressure in the clearance space. 
328. Size of Dry-air Pumps. — " Dry-air" pumps are used in con- 
nection with barometric and surface condensers where a high degree 
of vacuum is essential, as in steam turbine practice. Such pumps are 
intended to exhaust the saturated non-condensable vapors only. 



628 STEAM POWER PLANT ENGINEERING 

The capacity of the dry-air pump is based upon experience rather 
than theory. Current practice gives 

Q = 20 v to 30 v for vacua under 27 inches. 

Q — 35 v to 50 v for vacua of 28 inches and over, both referred to a 
30-inch barometer. 

Professor Weigh ton states that "with suitable condenser arrange- 
ments and a reasonably air-tight system there is nothing gained in 
efficiency by the use of air pumps exceeding in capacity 0.7 of a cubic 
foot per pound of steam condensed up to a vacuum of 29 inches. " 
(Engineering Record, May 19, 1906, p. 61.) 

The work done by the average "high-vacuum" reciprocating dry- 
air pump is a maximum for vacua between 18 and 20 inches. 

This may be proved from Fig. 406 which represents a theoretical in- 
dicator card from the air-pump cylinder. 

Let p 2 = pressure in the condenser, pounds per square inch absolute; 
px = atmospheric pressure; 

v 2 = piston displacement, including clearance, cubic feet; 
Vi = volume of air in the cylinder when the valve opens to at- 
mosphere, cubic feet; 
v c = clearance volume, cubic feet. 

The work done is proportional to the area ABCD. 
Area, ABCD = work done = area EBIO + BAGI - FAGO - ECDF. 
Neglecting the exponential factor n for the sake of simplicity, thus 
making pv = p±vi = p 2 v 2 = constant, 

P2V2 — PiVc + PiVc (244) 

Substitute p\V\ for its equivalent P2V2 and ¥LLl for its equivalent p 2 v c 
and integrate. 

W = piVi + piDx log e v 2 - P1V1 log e vi - pm + 2i-Mi (245) 

v 2 

making the first derivative zero. 

^ = = log e vt-1- log c vi + -> (246) 

dv v 2 

(247) 

(248) 
p 2 v 2 . (249) 









= 


log 


V2 _ 
°'Vi 


-1 + 


Vc 
— ) 

v 2 


i.e., 


W 


is 


a maximum when 
-log e - = 


Vc _ 

v 2 


-1, 






or 






log e ^ = 

V2 


1 - 


v 2 ' 


since 


PlVi 






PUMPS 



629 



For average high-vacuum practice v c = 3 per cent of the piston dis- 
placement. Assume v 2 — 1, v c = 0.03, and pi = 14.7 pounds per square 
inch, and substitute these values in Equation (249), thus: 

, 14.7 , 0.03 

loge ^7 = 1 -nro' 

Whence p 2 = 5.5 pounds per square inch absolute, which corresponds 
to a vacuum of 18.6 inches of mercury. 




Fig. 406, 



Thus we see that the maximum load on the pump occurs when the 
vacuum is between 18 and 20 inches. If the vacuum is less than this, 
the load falls off because of the decreased difference in pressure. If the 
vacuum is greater, the load falls from the decrease in weight of air 
handled. 

329. Centrifugal Pumps. — Centrifugal pumps consist of two essen- 
tial elements, (1) a rotary impeller which draws in the water at its 
center and (2) a stationary casing which guides the water thrown from 
the ends of the impeller to the discharge outlet. Increase of peripheral 
speed increases the energy in the impeller. This increase in energy may 
take the form of increase in pressure or potential energy, or it may be 
in the form of increase in rate of flow or kinetic energy. In general 
there is an increase in both kinetic and potential energy. The im- 
peller may be of the open type, Fig. 408 (B), or closed, Fig. 408 (A). 
The casing may be cylindrical and concentric with the impeller, Fig. 
412, or of spiral form, Fig. 407. It may be plain or fitted with diffu- 
sion vanes and any number of impellers may be employed. The shape 
of the impeller and casing and the number of impellers or stages deter- 
mine the efficiency of the pump and its adaptability to certain con- 
ditions of service. 



630 



STEAM POWER PLANT ENGINEERING 



Discharge 



Centrifugal pumps are generally classified as 

1. Volute 

2. Turbine. 

330. Volute Pumps. — Fig. 407 gives an end view of a typical single- 
stage volute pump with end plate removed so as to expose the im- 
peller, and Fig. 409 shows a section through a modern single-stage 

volute pump with double suction. In the 
volute pump the casing is of spiral design 
forming a gradually increasing water or 
" whirlpool" chamber, A-B, Fig. 409, for 
the purpose of partially converting velocity 
head to pressure head. The older forms of 
volute pumps were very inefficient, seldom 
delivering more than 40 per cent of the 
energy supplied and usually not adapted 
to lifts greater than 50 feet. The modern 
pumps give efficiencies as high as 80 per 
cent, and the lift is limited only by the speed 
of the impeller. As a general rule the volute 
pump is of single-stage construction and limited to comparatively low 
lifts, 120 feet and under, though two-stage pumps of this type are on 
the market designed for heads as high as 1000 feet. 




Fig. 407. 



Suction 
A Typical Centrifugal 
Pump. 





Fig. 408. Types of Impellers. 

331. Turbine Pumps. — In the usual design of volute pumps the 
stream of water in the casing is at cross current with that thrown out 
from the impeller as shown in Fig. 410. The turbine pump is provided 
with a system of diffusion vanes or expanding ducts, disposed between 
the periphery of the impeller and the annular casing, somewhat like 
the guide vanes in a reaction turbine water wheel, so that the fluid 
emerges tangentially at about the velocity in the casing (see Fig. 411). 
The casing is usually concentric with the impeller and of uniform cross 
section though the volute casing is sometimes used in this connection. 
For high lifts these pumps are compounded, thereby reducing the 






PUMPS 



631 



peripheral velocity and decreasing the friction losses. Fig. 412 shows 
a section through a three-stage Worthington turbine pump as installed 




Fig. 409. Typical Single-stage Double Suction Volute Pump. 

in the testing laboratories of the Armour Institute of Technology and 
designed to deliver 200 gallons per minute against a 750-foot head at 
2500 r.p.m. 





Fig. 410. Direction of Water from the 
Impellers of a Centrifugal Pump 
without Diffusion Vanes. 



Fig. 



411. Effect of Diffusion Vanes on 
the Direction of Water. 



332. Field for Centrifugal Pumps. — In view of past developments 
it is probable that the centriugal pump will supplant the piston type 
of pump for practically all purposes, except perhaps for deep-well service 
and for very heavy pressures. Centrifugal pumps are now used for 
boiler feeding, circulating condensing water, hot-well and wet-vacuum 
purposes and for various applications of industrial service. Efficiencies 



632 



STEAM POWER PLANT ENGINEERING 



above 70 per cent are not unusual and the head against which the 
pump may operate is limited only by the peripheral speed at which 
the impeller may be safely run. Although the equivalent heat efficiency 
of the high-grade piston pump is superior to that of the centrifugal 
pump, other items, such as low first cost, decreased cost of repairs and 
the like, frequently offset this advantage. Some of the advantages of 
the centrifugal pump as compared with the piston type are : 

1. Low first cost, 

2. Compactness, 

3. Absence of valves and pistons, 

4. Low rate of depreciation, 




Fig. 412. Worthington Three-stage Turbine Pump. 



5. Uniform pressure and flow of water, 

6. Simplicity of design and ease of operation, 

7. Freedom from shock, 

8. High rotative speed, permitting direct connection to electric 

motors and steam turbines, 

9. Ability to handle dirty water, sewage and the like, 

10. In case of stoppage of delivery, the pressure cannot increase 

beyond the predetermined working pressure, and 

11. Ease of repair. 

Some of the disadvantages are: 

1. Efficiency not as high as the best grade of piston pumps, 

2. Cannot be direct connected to low-speed engines when high lifts 

are desired, and 

3. The rate of flow cannot be efficiently regulated for wide ranges 

in duty. 



,160 


























































































































140 










































\ 












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^ 


tf* 


<0> 










































120 












0^ 












































































































100 


























































































































80 


































































































































% 


60 




















>YVC 


L* 






















) 
































J& 










































40 


























































































































20 
































































/ 




























































/ 





























































20 



40 



60 80 100 120 

Capacity 



140 



Fig. 413. Centrifugal Pump Characteristic for Hydraulic Elevator Service, Boiler 

Feeding, etc. 



























































































































120 
























































































































100 




























































% m 


































Ef 


fici 


eirc 






















































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o 

W 60 




























c£ 






























































>te 


*fi 


























40 








































^ 


^s 




























































V 


V 






\ 








20 
















































s 




\ 
























































\ 


\ 




























































\ 










20 40 60 80 100 120 140 
Capacity 
Fig. 414. Centrifugal Pump Characteristic for Dry-Dock Service. 


























































































































140 






















































































































120 






















































































































rcjlOO 
















Dh 


ira 


ate 


is: 


\c 




































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© 

W 80 










































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—= 






























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fcfl^ 














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40 


































































S 


S 


















































20 






/* 


' 






















































/ 


























































Z. 


1 
















































1 





40 



120 



140 



60 80 100 

Capacity 
Fig. 415. Centrifugal Pump Characteristic for Water Works with Large Friction Head- 



634 



STEAM POWER PLANT ENGINEERING 



6U 






































70 










































Cap 


and I 


: ea a 




























60 










































































50 
















































\/ 




































A 


' / 


— o,<? 


L* 
































~*fi 




%j 


























BO 










































































20 




s 1 
























































SP 


EED 


e.90 


R.P 


M.. 






10 




1 


































/ 




































°. 


/ 




































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10 


30 






20 


00 






30 


00 






4000 





7.0 



50 ! 



40 



u 
o 
30 pu 

>> 

o 

© 



10 



Gallons per Minute 



>k130 

1 12 ° 
I no 

100 



90 



50 




LEA-DEGEN .CENTRIFUGAL PUMP 

10 INCH TWO STAGE MOTOR DRIVEN 
DESIGNED for 
GAL. P.M. X '100 FEET 8 AT 600 R. P r M. 



200 400 600 800 1000120014001600180020002200240026002800300032003400360038004000 
Gallons per Minute 

Fig. 417. Performance of Two-stage Lea-Degen Centrifugal Pump. 



PUMPS 



635 



333. Performance of Centrifugal Pumps. — For best efficiency a cen- 
trifugal pump must be properly designed for the intended service as 
to curvature of vanes, diameter and speed of impeller, and number of 
stages. Figs. 413 to 415 are based upon experiments with De Laval 
centrifugal pumps. When a practically uniform head is required at 
constant speed with varying water supply as in city water works, hydrau- 
lic elevator systems or boiler feeding, the impeller vanes are designed 
to give the characteristic curve illustrated in Fig. 413 which protects 
the motor from possible overload. 




800 1000 1200 
Gallons per Minute 
Fig. 418. Characteristic Curves of 8-inch Kerr Centrifugal Pump for Low Heads 
and Steam Turbine Speeds. 

In dry-dock and other variable-head work, in order not to overload 
the motor, the power should be practically constant through wide 
variations of head and at the same time the efficiency should not vary 
seriously. A desirable characteristic for such a pump is illustrated 
in Fig. 414. 

In water-supply systems in which the friction of the piping is a large 
part of the total head at full delivery, the characteristic shown in Fig. 415 
is especially useful. Thus, when the system reduces its demand for 
water and the frictional head is consequently considerably reduced, 
the pump would automatically adjust itself to the reduced head without 
change of speed. Figs. 416 to 419 are based upon experiment and show 
the relationship between speed, head, capacity, efficiency and power 
consumption of various types of pumps. 

The curves in Fig. 418 are of interest in that they show the charac- 
teristic of a centrifugal pump operating at low heads for very high 
rotative speed. 



636 



STEAM POWER PLANT ENGINEERING 



8 



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K 

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PUMPS 



637 



Tables 102 to 104 give the capacity, speed, head and power require- 
ments for commercial sizes of centrifugal pumps, and may be used as 
a guide in selecting the size of pump for general service. 







Pump Designed for 250 Gall.E M.x 

700 Feet at 2000 and 20000 R.P.M. 

Tested April 8th & 9th, 1004 




1000 
900 


1 

i 














1 




I 


















I 


1 


2iai^ 


**a. 














P 


■§700 

a 

3 




\0, 

\*5 






^"^s 


\ 




/ 








\ 


m 


W///W 

1 


\ 




1 






\ 


k 




f 




y 






400 
300 
200 
100 
0- 






J* 


gfr 


.-—-*• 


I- — *x 










4 


P' 








e\l 






4 
1 


f 













\ 




<3/ 














V 




/ 














1 

! 


1-1 



150 200 250 300 
Gallons per Minute 

Fig. 419. 



350 400 



no 



-i 

Sk 

50 t 
o 



30 S 



10 



TABLE 103. 
DATA PERTAINING TO WORTHINGTON MULTI-STAGE TURBINE PUMPS. 



t& 


a 
o 

o . 

CQ QJ 

-•§ 

o c 

is 

5 

1.5 

2 

2.5 

3 

4 

5 

6 

7 
10 
12 


"3 „? 

a* 

si 

el 


K 

IS 

P4 o 


Total Head in Feet, R.P.M., Number of Stages. 


5.- 


100 Ft. 


200 Ft. 


300 Ft. 


400 Ft. 


Diameter c 
charge P 
Inches. 


Ah' 


o . 

2 
2 
2 
2 
2 
2 
1 
1 
1 
1 




S-i CD 

525 




o . 


P4 


o . 

|CQ 


I 


30 

45-60 

75-100 

125-150 

200-250 

350-450 

600-700 

800-1000 

1500-1800 

2500-2800 


0.02 

0.0395 

0.0625 

0.095 

0.134 

0.222 

0.297 

0.396 

0.643 

1.00 


2000 

1500 

1300 

1200 

1100 

950 

800 

750 

600 

500 












1.5 


2000 
1800 
1600 
1400 
1200 
1300 
1200 
1000 
800 


3 
3 
3 
3 
3 
2 
2 
2 
2 










2 


1500 
1300 
1200 
1000 
1150 
1000 
800 
700 


4 
4 
4 
4 
3 
3 
3 
3 






2.5 
3 
4 
5 
6 
8 
10 


1050 
950 
780 
670 


5 
5 
5 
4 
4 
4 
4 



Horse power based on maximum capacity. 



638 STEAM POWER PLANT ENGINEERING 

TABLE 104. 

CAPACITIES, HEADS AND SPEEDS OF McEWEN BROS. PUMPS 

Double Suction Hot-well Pumps. 





Speed, 
R.p.m. 


Range of 


Size. 


Economical Capaci- 
ties, Gal. per Min. 


Economical 
Head. 


Brake H.p. 


Pipe Velocities, 
Ft. per Sec. 


In. 
3 

4 
5 
6 


3600 
3200 
2800 

3600 
3200 
2800 

3600 
3200 
2800 

3600 
3200 
2800 


125- 160 
110- 145 
100- 120 

250- 325 
225- 300 
200- 250 

450- 600 
400- 550 
350- 500 

700-1000 
650- 900 
600- 800 


Ft. 

60-50 
45-35 
35-30 

60-50 
45-35 
35-30 

85-70 
65-55 
50-40 

85-70 
65-55 
50-40 


4.5- 4.7 
3.2- 3.5 
2.4- 2.5 

8.1- 8.75 
5.7- 6.2 
4.1- 4.3 

20 -22 

14 -16 

9.7-11 

30 -36 
23 -27 

16 -18 


5.7- 7.2 

5.0- 6.1 

4.5- 5.4 

6.4- 8.3 

5.8- 7.7 

5.1- 6.4 

7.4- 9.8 

6.6- 9.0 

5.8- 8.2 

8 -11.4 
7.4-10.2 
6.8- 9.1 



Turbine Pumps: Cassier's Mag., April 12, 1911; Engng., Jan. 26, 1912; Eng. 
Rec, June 15, 1912. 

334. Rotary Pumps. — Rotary pumps are often used for circulating 
cooling water in condenser installations, and give about the same effi- 
ciency as centrifugal pumps under similar conditions of operation. 
For moderate pressure and large volumes they offer the advantage of 
low rotative speed, thus permitting direct connection to slow-speed 
steam engines. At high speeds they are noisy, due chiefly to the gear- 
ing. They occupy considerably less space than piston pumps of the 
same capacity, but require more room than the centrifugal type. 

Fig. 420 shows a section through a two-lobe cycloidal pump. The 
shafts are connected by wheel gearing, the power being applied to one 
of the shafts. The water is drawn in at I and forced out at 0, the 
displacement per revolution being equal to four times the volume of 
chamber A. There is no rubbing between impellers and casing. In 
this type of pump the pressure is independent of the speed of rotation, 
and the capacity varies almost directly with the speed. The slip varies 
from 5 to 20 per cent according to the discharge pressure. 

Fig. 421 shows a section through a rotary pump with movable but- 
ment. Fig. 422 illustrates the performance of a 45-mm. Siemens- 
Schuckert rotary pump at different speeds and discharge pressures. 
(Zeit. d. Ver. Deut. Ing., June 24, 1905, p. 1040.) Large rotary pumps 



PUMPS 




639 



Fig. 420. Two-Lobe Cycloidal 
Pump. 




Fig. 421. Rotary Pump with 
Movable Butment. 



a 


65 


O 








<D 




Pn 




$ 


CD 


§60£ 


o 


<x> 


& 


a 


calE 


o 


a 


55 


| 




i 




X 





2.0 



1.6 

1.4 
1.2 
1.0 

0.8 
0.6 
0.4 
50 0.2 























Efficien 


cy 


^^ 




^, — 


































}^^ 










^V0?\ 


\$^^ 




_ 










^ots^ 


tf!&^^ 










O^ 








^ 








Average I 


lead, 75 Ft. 



















400 



&70 

c 



o 2.0 



1 65 I 1.8 



W 



1.2 



1.0 



500 



600 700 

Revolutions Per Minute 



900 



Head Constant, Speed Variable. 



50 



45 



40 



25 ; 



20 



1000 

























. 




















"^ 










-t?ffi< 


\<t&2~ 


























?o<0* 














. 








$& 






















^ 








Ztffi 


\ig&~ 





















^> 




Average Speed 812 R.P.M. 
Average Capacity 
45.5 Gal.Per Minute 


•^ 







































70 80 



90 



100 110 120 130 140 150 160 170 180 
Total Head, Eeet, _ 



Speed Constant, Head Variable. 
Fig. 422. Performance of a Small Rotary Pump. 



640 



STEAM POWER PLANT ENGINEERING 



give much higher efficiencies, but the general characteristics are about 
the same. A combined efficiency of pump and engine as high as 84 per 

cent has been recorded. 
(Trans. A.S.M.E., Vol. 
24, p. 385.) 

335. Circulating Pumps. 
— This term is ordinarily 
applied to the pumps 
which supply injection 
water to the condenser. 

Three types are found 
in practice: the piston, 
the centrifugal, and the 
rotary pump. Figs. 296 
and 309 show the appli- 
cation of reciprocating 
pumps to condenser in- 
stallations and Figs. 312 
and 313 a similar applica- 
tion of centrifugal pumps. 

For large volumes of 
water and low heads the 
centrifugal or rotary 
pump is generally adopt- 
ed on account of mini- 
mum space requirements 
and low first cost. 

In very large central 
stations where the de- 
mand for circulating 
water is enormous and the 
lift is moderately high, 
the high-duty pumping 
engine is sometimes in- 
stalled. Fig. 423" shows 
a section through one of 
the nine high-duty circu- 
lating pumps at the New 
York Rapid Transit 

Fig. 423. 10,000,000-Gallon Circulating Pump. Compan y» s powe r house. 

The steam end is operated by Corliss cylinders and is of the cross- 
compound type. The maximum capacity is 10,000,000 gallons per day 




PUMPS 



641 



• ALL VALVE 




OISCHARGE 
*S fv V«LVE 



(24 hours) against a head of 50 feet at mean low water. The actual 
lift is much less than this, as the discharge is aided by the vacuum in 
the condenser. 

336. Hot-well Pumps. — In surface-condenser practice the condensed 
steam is often handled by a small independently driven pump called 
the hot-well pump. The piston type of pump is being rapidly superseded 
by the turbine-driven centrifugal bt E »m 

pump in this connection. The water 
from the condenser hot-well should 
flow to the pump under a head of 
3 or 4 feet or more, if convenient. 
If the head on the suction side is 
less than this the pump "cavitates" 
or becomes vapor bound and is 
unable to remove the water. The 
discharge over the pump should 
be provided with a check valve 
to prevent water returning to the 
condenser. Centrifugal hot-well 
pumps are ordinarily designed with 
horizontal suction and vertical dis- 
charge to minimize air pockets in 
the volute. These pumps are ordi- 
narily operated without automatic 
control and are permitted to operate 
at constant speed. 

337. Air Lift. — The air lift is a 
simple arrangement of piping where- 
by water may be raised by means of 
compressed air. There are no work- 
ing parts, and no valves are em- 
ployed except to regulate the supply FlG - 424 - The Pulsometer. 

of air. Its particular field of application lies in pumping water from 
a number of scattered wells, and on account of the total absence of 
working parts it is peculiarly adapted to handling water containing 
sand, grit and the like. The device consists of a partially submerged 
water pipe and air supply variously arranged as in Fig. 425 (A) to (D). 
Compressed air forced into the water pipe at or near the bottom de- 
creases the density of the column and the difference in weight between 
the solid column of water B and the air-water column A causes the flow. 
The successful operation of this device depends upon the ratio of the 
depth of submersion B to the total head C. 










642 



STEAM POWER PLANT ENGINEERING 



The quantity of air necessary to operate an air lift may be closely 
approximated from the equation (see Prac. Engr. U. S., April 1, 1912, 
p. 354) 

V = TT^TTZ , (250) 



in which 



log 34 X C 



V = cubic feet of free air per gallon, 

S = actual submergence in feet, 

C = coefficient determined from experiment. 




77/ "Wa ter LereL 
in Well 



Fig. 425. Different Arrangement of the "Air Lift." 

The actual submergence S may be determined from the relationship 

LSr 



in which 



o — — j — 1 

I'D 



L = actual lift in feet (A, Fig. 425), 

S p = submergence, percentage (100 ^, Fig. 425), 

"' A 

l p = lift, percentage (100-^, Fig. 425). 



(251) 



The coefficient C may be approximated as follows: 

C = 255 -0.1L. 



(252) 



For the air pressure required for any lift and any percentage of sub- 
mergence it is convenient to divide the actual submergence in feet by 
2 to get the gauge pressure in pounds. This gives enough pressure in 



PUMPS 



643 



excess of that due to water head to allow for the pipe friction and other 
losses. 

The efficiency (" water" horse power divided by "air" horse power) 

varies from 30 to 50 per cent, increasing as the ratio -^ increases from 

0.55 to 0.85. (Engineer, U. S., Aug. 15, 1904, p. 564.) A number of 
tests gives efficiencies (" water" horse power divided by i.h.p. of steam 
cylinder) varying from 20 to 40 per cent. The horse power required 
to compress one cubic foot of free air to different pressures per square 
inch, as determined from actual practice, is approximately as follows: 



Pressure in 
Pounds. 


Horse Power 
Required to 

Compress 
1 Cubic Foot. 


Pressure in 
Pounds. 


Horse Power 
Required to 

Compress 
1 Cubic Foot. 


176 
140 
100 

80 


0.434 
0.376 
0.201 
0.189 


60 
45 
30 


0.159 
0.145 
0.121 



(Engr., Lond., Aug. 14, 1903, p. 174, Dec. 11, 1903, p. 568, Feb. 12, 1904, p. 172.) 

When it becomes necessary to raise water to a height exceeding say 
175 feet above the level in the well, it is customary to use two or more 
pumps, the total lift being divided between them. 

Air Lift: Eng. and Contr., Feb. 14, 1912; Bulletin No. 450, Univ. of Wis.; Prac. 
Engr., April 1, 1912; Power, May 21, 1912. 

Pulsometer: Tech. Quar., Sept., 1901; Public Works, Aug. 15, 1904; Engr. U. S., 
July 15, 1904; Experimental Eng., Carpenter, p. 621; Thermodynamics, Wood, 
p. 293; Trans. A.S.M.E., 13-211. 

Cost of Operating American Pumping Stations: Eng. Rec, Aug. 6, 1904; Proc. 
Engrs. Club of Phil., Oct., 1906. 

Complete Description of Various American Types of Steam, Rotary, and Centrifugal 
Pumps: Engr. U. S., Jan. 1, 1904. 

The Selection of Waterworks Pumping Machinery: Eng. and Contr., April 24, 
1912. 

Centrifugal Pump for Boiler Feeding: Power and Engr., Mar. 15, 1910. 

Recent Records of High Duty Pumping Engine: Eng. News, Feb. 3, 1910. 

The Hydraulic Ram: Engr. and Contr., Jan. 3, 1912; Mar. 22, 1911. 



CHAPTER XIV. 

SEPARATORS, TRAPS, DRAINS. 

338. Live-steam Separators. General. — The function of a steam sepa- 
rator is the removal of entrained water from steam. 

Unless a boiler is liberally provided with superheating surface, the 
steam may contain an amount of moisture varying from 0.3 to 5 per 
cent. If the boiler is poorly proportioned or forced far above its rating, 
this percentage may be greatly increased. The quality of the steam 
is still further reduced by condensation in the steam pipe, which may 
vary from 1 to 10 per cent, depending upon the length of pipe and 
efficiency of covering. 

One of the effects of moisture in steam is to increase its density and 
reduce its elastic force. It also increases its conductivity, so that 
during the work of expansion more heat is absorbed from the walls of 
the cylinder and discharged into the atmosphere or into the condenser 
without doing useful work. (Ewing, "The Steam Engine," p. 151.) 
Although the heat loss from this cause is small, the danger arising from 
the introduction of a considerable amount of water in the cylinder 
renders the removal of the moisture necessary. See par. 193 for in- 
fluence of moisture on steam consumption. 

The essentials of a good separator are high efficiency as a water 
eliminator, ample storage capacity for any sudden influx of water, 
simplicity and durability in construction, and small resistance to the 
current of steam passing through. A good separator may be relied 
upon to remove practically all of the moisture from steam containing 
under ten per cent entrainment and all but two per cent from steam 
containing as much as twenty per cent. (Engineer, U. S., Jan. 15, 
1904.) 

Table 105 gives the results of a series of tests made by Professor R. C. 
Carpenter in 1891 of six steam separators. (Power, July, 1891, p. 9.) 
Conclusions from these tests were: 

1. That no relation existed between the volume of the several sepa- 
rators and their efficiency. 

2. No marked decrease in pressure was shown by any of the separa- 
tors, the most being 1.7 pounds by separator E. 

644 



SEPARATORS, TRAPS, DRAINS 



645 



3. Although changed direction, reduced velocity, and perhaps cen- 
trifugal force are necessary for good separation, still some means must 
be provided to lead the water out of the current of the steam. 

A series of tests made at Armour Institute of Technology in 1905 on 

a number of separators showed that the efficiency of separation decreased 

as the velocity of the steam increased* At the low velocity of 500 feet 

per minute all separators were equally efficient, at a velocity of 5000 

feet per minute several had little effect on eliminating the moisture 

present, and at a velocity of 8000 feet per minute only one gave efficient 

results. 

TABLE 105. 

TESTS OF STEAM SEPARATORS. 

(R. C. Carpenter.) 





Test with Steam of about 10 
Per Cent of Moisture. 


Make of 
Separator. 


Quality of 
Steam 
Before. 


Quality of 
Steam 

After. 


Efficiency. 


B 


Per Cent. 
87.0 
90.1 
89.6 
90.6 
88.4 
88.9 


Per Cent. 
98.8 
98.0 
95.8 
93.7 
90.2 
92.1 


Per Cent. 
90.8 
80.0 
59.6 
33.0 
15.5 
28.8 


A 


D 


C 


E 

F 





Tests with Varying Moisture. 



Quality of 
Steam Before. 



Per Cent. 

66.1-97.5 

51.9-98 

72.2-96.1 

67.1-96.8 

68.6-98.1 

70.4-97.7 



Quality of 

Steam After. 



Per Cent. 

97.8-99 

97.9-99. 

95.5-98. 

93.7-98. 

79.3-98. 

84.1-97. 



Average 
Efficiency 



Per Cent. 
87.6 
76. 
71 
63 
36 
28. 



339. Classification of Separators. — Separators are based on one or 
more of the following principles of action: 

1. Reverse current. The direction of the flow is abruptly changed, 
usually through 180 degrees. This causes the water in the steam, on 
account of its greater specific gravity, to be thrown into a receiving 
vessel, while the steam passes on in a reverse direction. 

2. Centrifugal force. A rotary motion is imparted to the steam 
whereby entrained water particles are eliminated by centrifugal force. 

3. Baffle plates. The flow is interrupted by corrugated or fluted 
plates to the surfaces of which the water particles adhere and from 
which they fall by gravity to the well below. 

4. Mesh. The separation is brought about by mechanical filtration 
through screens or meshes. 

The following outline shows the classification of typical separators, 
in accordance with the above principles: 

* See Power, May 11, 1909, p. 834. 



646 



STEAM POWER PLANT ENGINEERING 



Live-steam separators . 



Exhaust-steam separators 



Averse current \^3S^%%. 

f Keystone, Fig. 428. 

Centrifugal \ Mosher. 

[Robertson. 

f Bundy, Fig. 429. 

Baffle plate \ Austin, Fig. 430. 

( Detroit. 

Mesh (Direct, Fig. 431. 

iVlesn I Potter. 

( Jacketed baffle Baum, Fig. 432. 

( Absorption Loew, Fig. 433. 



340. Reverse-current Steam Separators. — Fig. 426 shows a section 
through a Hoppes steam separator and illustrates the principle of 

reverse-current separation. Steam may 
flow through in either direction. Both 
the inlet and outlet ports are surrounded 
by gutters C, C, partly filled with water, 
which intercept the moisture following 
the surface of the pipe, while the down- 
ward plunge of the steam throws the 
entrained water to the bottom of the 
separator. The condensation is carried 
from the troughs by pipe P to the well 
below, from which it is trapped at D in 
the usual way. The velocity of the steam 
in passing through this separator is greatly 
reduced to prevent the steam from taking 
up the water in the bottom of the well. 
This is brought about by increasing the 
area of the passage through the separator. 
Fig. 427 gives a sectional view of a Stratton separator, which, though 
primarily of the reverse-current type, embodies also the principle of 
centrifugal force. The separator consists of a vertical cast-iron cylinder 
with an internal central pipe C extending from the top downward for 
about half the height of the apparatus, leaving an annular space be- 
tween the two. The current of steam on entering is deflected by a 
curved partition and thrown tangentially to the annular space at the 
side, near the top of the apparatus. It is thus whirled around with all 
the velocity of influx, producing the centrifugal action which throws 
the particles of water against the outer cylinder. These adhere to the 
surface, so that the water runs down continuously in a thin sheet around 
the outer shell into the receptacle below. The steam, following in a 
spiral course to the bottom of the internal pipe, abruptly enters it, and 




Fig. 426. 



Hoppes Steam Sepa- 
rator. 



SEPARATORS, TRAPS, DRAINS 



647 



passes upward and out of the separator without having once crossed 
the stream of separated water. The rapid rotation of the current of 
steam imparts a whirling motion to the separated water which tends to 
interfere with its proper discharge from the apparatus. The separator 
has therefore been provided with wings or ribs E projecting at an acute 
angle to the course of the current, which have the effect of breaking up 
this whirling motion and allowing the water to settle quietly at the 
bottom, whence it passes off through the drain pipe D. 





Fig. 427. Stratton Steam Separator. 



Fig. 428. Keystone Steam Separator. 



341. Centrifugal Steam Separators. — Fig. 428 shows a section through 
a Keystone or Simpson's centrifugal separator. The separator consists 
of a cast-iron cylinder with vertical pipe C extending downward about 
two-thirds of the whole length; this pipe has a thread or screw wound 
spirally around it, the space between the threads being somewhat 
greater than the area of the steam pipe. The steam passing around 
the spiral course causes the water to be thrown against the outer walls 
by centrifugal force, while the dry steam passes through the small holes 
in the central pipe. The water passes down the outer walls, where its 
motion is arrested by obstructing ribs E, and is thence carried away by 
a drip pipe D to a suitable drain. 



648 



STEAM POWER PLANT ENGINEERING 




342. Baffle-plate Steam Separators. — Fig. 429 gives an interior 
view of a Bundy separator and illustrates the application of baffle 

plates for live-steam separation. This separator 
consists of a rectangular cast-iron casing with a 
cylindrical receiver beneath it. Directly across the 
steam passage are baffle plates corrugated for the 
reception of entrained water. The plates consist of 
vertical castings, each containing a main artery or 
channel which leads directly to the receiver. The 
fronts of the plates are flat, with a series of recesses 
sloping inwards and downwards, terminating in an 
opening of capillary size leading to the main artery. 
The plates are staggered, so that the steam must 
impinge against all of them in its passage. The 
particles of water adhere to the plates, collect, and 
Fig. 429. Bundy Steam f all by grav i ty mto t ^ e receiver. The flanges at the 

bottom constrict the opening of the reservoir so as 
to prevent the steam from picking up any portion of the water. 

Fig. 430 shows a section through an Austin separator and illustrates 
another class embodying the fluted baffle- 
plate principle. The steam in passing 
through the chamber impinges against the 
fluted baffle plate B. The moisture adheres 
to the surfaces, collects and trickles along 
the corrugations to the bottom of the well. 
These corrugations are formed in such a 
manner that the steam cannot come in con- 
tact with the water particles after they have 
been once eliminated. A perforated dia- 
phragm D prevents the water in the well 
from coming in contact with the steam. The 
current of steam is also reversed, thus giving 
additional separating properties to the appa- 
ratus. 

343. Mesh Separators. — Fig. 431 shows a 
section through a " direct" separator, illus- 
trating the principle of mesh separation. 
These separators are made with steel bodies 
and cast-iron heads and bases, in all sizes 
up to six inches inclusive, the larger sizes 
being constructed of cast iron or boiler plate. The cone C, perforated 
lining E, and diaphragm S are made of cold-rolled copper; the cone 




Fig. 430. 



Austin Steam Sepa- 
rator. 



SEPARATORS, TRAPS, DRAINS 



649 



is a substantial gray-iron casting, resting on three cast-iron supports 
hooked over the top of inner pipe as indicated. The method of operation 
is as follows: The accumulated moisture around the walls of the steam 
pipe is caught by the upper edge of cone C 
and carried down back of lining E to the 
water chamber. The current of steam 
entering the separator impinges upon the 
conical surface, which is composed of solid 
plate covered with sieve S, through which 
water may freely pass but from which it 
cannot readily escape. Passing through 
the sieve and depositing on the solid surface 
of the cone 0, this water is carried by con- 
ductors P to the water chamber. Perfo- 
rated lining E permits the moisture content 
of the steam to pass through the opening 
to the water below and prevents it from 
coming in contact again with the current 
of steam. A trough is provided at the 
lower edge of the inverted cup which leads 
all the water that may adhere to it to the 
water chamber. The steam flows through 
the passages indicated by arrows and is 
subjected to a whipsnapping action which 
tends to throw off any remaining mois- 
ture. The perforated plate D prevents 
the steam from picking water out of the 
water chamber. 
344. Location. — Live-steam separators may be located 

1. Inside the boiler, 

2. Between boiler and engine, 

3. At the steam chest. 

Where the steam pipe is very short, and particularly in marine and 
locomotive work where the tossing of the boiler induces excessive 
priming, the separator may be placed inside the boiler and its function 
becomes that of a dry pipe. In this location it prevents the water due 
to foaming and priming from passing to the engine, and reduces con- 
densation in the pipe by supplying dry steam. The " Potter mesh" 
and the "De Rycke centrifugal" are types of separators designed for 
this service. 

The arrangement of separator between engine and boiler, other than 
at the throttle or inside the boiler, is sometimes necessary for economy 




Fig. 431. "Direct" Steam 
Separator. 



650 STEAM POWER PLANT ENGINEERING 

of space. Where possible, however, the separator should be placed 
close to the steam chest. 

Current practice recommends that a receiver separator, which is an 
ordinary separator with a volume of two to four times that of the 
high-pressure cylinder, be placed close to the engine if the load is in- 
termittent or sharply fluctuating. This forms a cushion for absorb- 
ing the force of the blows caused by cut-off, delivers steam at a prac- 
tically uniform pressure, and reduces the vibration of the piping to a 
minimum. It also provides a reservoir for sudden demands made by 
the engine. Smaller pipes and higher velocities may be used with this 
arrangement. 

345. Exhaust-steam Separators and Oil Eliminators. — The function 
of an exhaust-steam separator is the removal of cylinder oil from the 
steam exhausted by engines and pumps. In plants where exhaust 
steam is used for heating it is quite essential to remove the oil from 
the steam before it enters the heating system, for the oil not only re- 
duces the efficiency of the radiators by coating them with an excellent 
non-conducting film but is an element of danger to the boiler itself. 
In condensing plants the separator will prevent the oil from fouling 
the condenser tubes and those of the vacuum heater if one is installed; 
this is an important factor, since the oil or grease lowers the efficiency 
of the heat transmission. 

In a general sense a live-steam separator is also an oil eliminator, and 
all the separators previously described perform this function to a cer- 
tain extent, since the underlying principles governing the elimination 
of oil from exhaust steam are similar to those employed in removing 
water from steam. Most of the separators described above are also 
designed in lighter form, as oil eliminators, but by far the greater 
number are based on the fluted baffle-plate principle, of which the Hine, 
Bundy, Cochrane, Utility, Peerless, and Keiley are well-known ex- 
amples. This type of oil separator will eliminate a considerable portion 
of the oil in the steam, provided the baffle plates or corrugated surfaces 
are frequently cleaned. 

The following is taken from the report of Professor R. Burnham of 
the Armour Institute of Technology on the test of a six-inch horizontal 
oil separator of the baffle-plate type: 

"For purposes of test the separator was placed in the exhaust line 
of a 9 X 18 X 24 cross-compound Corliss engine running under its maxi- 
mum load at 80 pounds pressure and exhausting into a Wheeler sur- 
face condenser against 26 inches vacuum. 

" Cylinder oil was fed through the lubricators of the high and low 
pressure cylinder at the rate of from 5 to 20 drops per minute, a record 



SEPARATORS, TRAPS, DRAINS 



651 



being made of the exact quantity of oil fed per hour. The separator 
was so arranged, by means of a receiver connected to the air pump, that 
the accumulation of oil and water could be readily trapped from it at 
any time. In order to determine the quantity of oil given up by the 
condenser, and not properly charged against the separator, each series 
of efficiency tests was preceded by a run of three hours during which 
time no oil whatever was fed to the cylinders. During the last hour a 
record was made of the weight of steam used and a sample of the con- 
denser discharge retained for analysis. 

"The efficiency tests were made by feeding at an excessive rate 
through the lubricators as described above, and when conditions became 
practically constant, records were made for one hour of the weight of 
oil used, weight of condensed steam, and drain from separator. Samples 
of the two latter were retained for analysis and the percentage of oil in 
them accurately determined, correction being made for the oil given up 
by the condenser. A second series of tests was made exhausting at 
atmospheric pressure. The results obtained are tabulated below. 



Oil in condensed steam with no oil feeding. 

(Charged to condenser.) Pounds per hour 

Oil fed to cylinder, pounds per hour 

Steam condensed per hour, pounds 

Oil caught by separator, per hour, pounds A 
Oil in condensed steam (corrected), pounds 

per hour B 

Percentage of oil in condensed steam by 

weight, per cent 

Efficiency of separator, percent , R ... 



Exhausting into 
26-inch Vacuum. 


iiixnausiing at 
Atmospheric 








Pressure. 


.051 


.057 


.0559 


.0353 


.0340 


.401 


.562 


.934 


.621 


.710 


1000 


1120 


1096 


905 


872 


.341 


.450 


.743 


.552 


.583 


.009 


.010 


.0096 


.0071 


.0050 


.0009 


.001 


.00088 


.00078 


.00057 


97.4 


97.8 


98.8 


98.7 


99.1 



"There was practically no free oil on the surface of the condenser 
discharge in any case, the small quantity of oil which passed the separa- 
tor (from 5 to 10 parts in a million of water by weight) existing as an 
emulsion, imparting a slight milky color to the water." 

It is a well-established fact that oil can be more effectually removed 
from wet than from dry steam, and some makers, notably the Austin 
Separator Company, inject a cold-water spray into the separator cham- 
ber. A similar result is brought about in the Baum separator, Fig. 432 y 
in which the corrugated baffle plate is hollow and cold water is forced 
through the chamber thus formed. Referring to Fig. 432 : The diverged 
baffle plate forms the wall of a chamber in which cold water is con- 



652 



STEAM POWER PLANT ENGINEERING 



FROM ENGINE 



tinually circulated. This circulation causes moisture to appear on the 
baffle-plate surface. The particles of oil, coming in contact with this 
moist surface as the steam current is diverged, adhere to it and fall 
by gravity into the well below, where they are completely isolated from 
the purified steam. A large portion of the oil and water, however, does 

not enter the separator at all but is caught 
by the inside ledge near the junction of 
the exhaust pipe and the separator. The 
oil and condensation which are carried 
along the bottom of the pipe come in 
contact with this ledge and are carried 
directly to the outlet pipe. 

A very successful method of removing 
oil from steam is to project the steam on 
to the surface of a body of water. The 
water may be hot or cold and will hold 
the oil if it once reaches the surface. It is 
essential, however, to reduce the velocity 
of the steam as it passes on its way to 
the outlet. Baldwin's grease separator is 
based upon this principle. (Baldwin on 
Heating, p. 234.) 

The most efficient method of removing 
oil is by combined filtration and absorp- 
tion. (Engineering News, May 22, 1902, 
p. 406.) A large chamber filled with 
coke, brick, broken tile, or other absorp- 
tion material is placed in series with the 
exhaust pipe. The steam passing through 
this chamber is entirely freed from oil and moisture, provided the ab- 
sorbing material is sufficient in quantity and is replenished as soon as it 
becomes saturated with oil. The annoyance attending the removal 
and replenishing of the absorbing material at frequent intervals and 
the great size of the apparatus are serious drawbacks. An example of 
this system of purification in which many of the objectionable features 
are reduced to a minimum is the Loew grease and oil extractor, Fig. 433. 
The exhaust steam enters the chamber at the top, strikes a large de- 
flecting plate shaped like an inverted V, and permits part of the con- 
densation and oil to be drawn off by the drain pipe. The steam then 
rises and is deflected, as indicated, against a series of shelves filled 
with fibrous material covered with coarse wire screens. The grease 
is removed from each shelf by suitable drains. This apparatus is 




Fig. 432. 



DRAJN 

Baum Oil Separator. 



SEPARATORS, TRAPS, DRAINS 



653 



sectional and any number of sections may be added without affecting 
the rest. 

In a non-condensing plant where the exhaust steam is used for heating 
purposes the oil separator is ordinarily placed in the main exhaust 
pipe just before it enters the heating system. Where several branches 
enter one main it is not customary to place a separator in each branch, 
one large separator located as above being sufficient. 

In condensing plants oil separators are 
seldom installed except where surface 
condensers are used, in which case the 
separator may be placed anywhere be- 
tween the engine and condenser. In case 
a vacuum heater is used the separator 
may be placed on either side of the heater, 
depending upon the type of separator. 
If the separator is of the " jacket-cooling" 
or " spray" type, it may be placed be- 
tween the engine and the vacuum heater; 
if, however, it is of the " baffle-plate" 
type, the oil will be more efficiently re- 
moved if the separator is placed between 
the heater and condenser so that it will 
get the benefit of the moisture formed in 
the heater. In the latter location, how- 
ever, the separator will not prevent the 
oil from fouling the heater tubes. 

Where a jet condenser is used and water is taken from the hot well, 
the hot well itself acts as an oil separator. (Trans. A.S.M.E., 24- 
1144.) 

All separators, steam and oil, should be provided with gauge glasses 
and should be thoroughly drained and the drainage should be auto- 
matic. 

346. Exhaust Heads. — The function of the exhaust head is the elimi- 
nation of oil and water from steam exhaust before permitting it to be 
discharged into the atmosphere. Unless removed, the water and oil 
rot the roofs and walls in summer and pollute the atmosphere sur- 
rounding the plant. The exhaust head also acts as a muffler reducing 
the noise of the escaping steam. Exhaust heads are built on the same 
principle as steam and oil separators and most separator builders manu- 
facture them. Fig. 434 shows a section through a typical exhaust head. 
The condensation is ordinarily drained to waste, though with proper 
purification it may be returned to the boiler. With an efficient oil sepa- 




Fig. 433. Loew Grease 

Extractor. 



654 



STEAM POWER PLANT ENGINEERING 



rator in the exhaust line the condensation in the exhaust head may be 

returned directly to the boiler without 
further purification. 

Live-steam separators are propor- 
tioned so that it is only necessary, 
in the average installation, to specify 
the size of pipe, the type of engine, 
the steam pressure, and the style, 
whether horizontal or vertical. 
Gauge glasses, gauge cocks, and com- 
panion flanges are usually provided 
by the maker. In some cases the 
capacity of the reservoir is also speci- 
fied. In specifying oil extractors the 
following additional data are neces- 
sary for an intelligent choice: the 
number of engines and pumps ex- 
hausting into the line, the location 
of the separator, the steam pressure, 
velocity, and the quality and quantity 
of cylinder oil used. A guarantee of 
efficiency and of material and work- 

Fig. 434. A Typical Exhaust Head. manship is often demanded. 




REFERENCES. 

Water and Oil Separators: Am. Elecn., Jan., 1900. 

Steam Separators: Am. Elecn., June, 1905. 

"Dry Steam": Goubert Mfg. Co., 85 Liberty St., N. Y. (Catalogue). 

Cochrane Separator: Harrison Safety Boiler Works, Philadelphia, Pa. (Catalogue). 

Bundy Separator: A. A. Griffin Iron Co., N. Y. (Catalogue). 

A Bad Case of Discharge Water with Steam from Water-Tube Boilers: Trans. 
A.S.M.E., Vol. 26. 

Location of Steam Separators: Power, Oct., 1904. 

Condensing Exhaust Head: Eng. News, Vol. 49, p. 419; Eng. Rec, Vol. 40, 
p. 177. 

Experiments in Separating Oil from Condensed Steam: Eng. News, May 22, 1903; 
Eng. Rec, April 27, 1901; Engr., Lond., March 12, 1897, Oct. 20, 1905; Power, 
Aug., Sept., 1896, May, 1903; Heating and Ventilation, Feb., 1897; Trans. A.S.M.E., 
Vol. 17, p. 295. 

Oil Separators: Baldwin on Heating, pp. 233-237. 

Oil Separation in a Combination Engine and Turbine Plant: Power, Oct., 1906. 

Test of a Cochrane Steam Separator: Power, April, 1898. 

Test of Lippincott Separator: Engr. U. S., Aug., 1902. 

Test of a Linstrum Steam Separator: Engr. U. S.., June 15, 1904. 

Test of an Austin Steam and Oil Separator: Trans. A.S.M.E., Vol. 20, p. 489. 






SEPARATORS, TRAPS, DRAINS 655 

Test of a Detroit Live Steam Separator: Engr. U. S., April 15, 1904; Power, Jan., 
1902. 

Tests of Direct Separators, Oil and Steam: Direct Separator Co., Syracuse, N. Y. 
(Catalogue) . 

Tests of Six Steam Separators: Power, July, 1891, p. 9; Engr. News, Vol. 26, 
p. 233. 

Tests of a "Utility" Oil Eliminator: Engr. Rec, May 2, 1903; Engr. Rev., May, 
1903. 

The Hot Well as an Oil Extractor: Trans. A.S.M.E., 24-1144. 

347. Drips. — No matter how thoroughly a steam pipe or reservoir 
may be covered with insulating material considerable condensation 
takes place. With the best covering this loss approximates one sixth 
of a pound of steam per square foot of pipe surface per hour for steam 
pressures of one hundred pounds, and runs as high as one pound of 
steam for bare pipes. See Table 113 for results of experiments on the 
loss of heat from bare pipes, and Table 114 for data on the efficiency 
of pipe coverings. In addition to this water of condensation, from \ 
to 2 per cent of moisture is carried over by the steam from the boiler. 
This water, unless thoroughly removed, is a constant source of danger 
to the engines and causes water hammer and leaky joints in the piping. 

A joint on a steam pipe may safely withstand a steam pressure of 
100 pounds without leaking and still leak badly under a water pressure 
of half that amount. This is due to the fact that the steam with its 
high temperature causes the pipe to expand, thus insuring a tight 
joint, while the entrained water (which cools as it collects) causes the 
pipe to contract and allows a leak. 

The entrained water and water of condensation are usually spoken of 
as " drips." Drips may be divided into two classes, low pressure and 
high pressure. 

348. Low-pressure Drips. — Low-pressure drips include the steam 
condensed in heating systems, exhaust steam feed heaters of the close 
type, exhaust steam piping, receiver barrels, steam chests, and exhaust 
heads. As these drips are impregnated with oil and are useless for 
boiler feed without purification, they are usually discharged to waste. 
Most city ordinances require the drips to be cooled to 100 degrees F. 
before being discharged into the sewer. In this case they must be 
first discharged into a tank and permitted to cool. This tank must be 
vented to the atmosphere to prevent back pressure. Fig. 435 shows 
an installation in which the heat abstracted from the drips, etc., is 
used to heat the feed water. The drips from the throttle valve and 
steam chest in a non-condensing plant are ordinarily discharged into 
the exhaust pipe as shown in Fig. 436. In a condensing plant the 
throttle drips are piped to a trap or to the free exhaust pipe. The re- 



656 



STEAM POWER PLANT ENGINEERING 



turns from a steam-heating system are sometimes classified as low- 
pressure drips. They are invariably returned to the boiler. 




Fig. 435. Closed Heater Installation for Abstracting Heat from Oily Drips. 

In small plants all the low-pressure drips may be connected to one 
large pipe and this pipe in turn to a single trap, provided there is but 

little difference in pressure in the various 
drip pipes. In case of different pressures 
separate leads should be run to waste or 
traps. 

The drips from the receiver and 
vacuum heater barrels in a condensing 
plant are oftentimes under less than 
atmospheric pressure, and sometimes 
the pressure varies from a slight vacuum 
to 10 or 20 pounds gauge, and conse- 
quently cannot be disposed of as de- 
scribed above. If possible, the heaters 
and receivers should be placed so as to 
drain into the condenser (see Fig. 449). 
Should this arrangement prove imprac- 
ticable, the barrels may be drained by a 
trap especially arranged as shown in Fig. 450. 

349. Size of Pipe for Low-pressure Drips. — In the average exhaust- 
steam feed-water heater one pound of steam in condensing gives up 




Fig. 436. Simple Method of 
Draining Drips. 



SEPARATORS, TRAPS, DRAINS 657 

approximately 1000 heat units. This will heat about 6 pounds of water 
from 60 to 200 degrees F. Hence the area of the drip which carries 
the water of condensation from the closed heater need be but one 
fifth that of the feed pipe. In no case, however, should a pipe smaller 
than one half inch in diameter be used. Should the same pipe be used 
for both exhaust head and heater drips, an area of one fourth area of 
feed pipe would prove of ample capacity. In practice it is customary 
to use the size of pipe conforming with the outlet furnished by the 
manufacturer of the apparatus, and' only when several pieces of ap- 
paratus are connected to one main are calculations made for the size 
of this main. 

The drip pipe from the throttle valve is ordinarily one half inch in 
diameter irrespective of the size of steam pipe; this is also true of the 
steam-chest drip. 

350. High-pressure Drips. — High-pressure drips consist of • those 
which are condensed under practically boiler pressure and include the 
steam condensed in steam pipes, cylinder jackets of engines, reheating 
coils of receivers, and separators. Being free from oil and containing 
considerable heat, they are usually returned to the boiler. Drips may 
be returned to the boiler automatically by means of 

1. Steam traps. 

2. Holly steam loop. 

3. Pumps. 

351. Classification of Steam Traps. — Steam traps may be divided 
into two classes, depending on their use, — return and non-return. 
Both of these two classes may be subdivided into five types according 
to the principle of operation, viz. : 

I. Float. III. Bowl. 

II. Bucket. IV. Expansion. 

V. Differential. 

Return Trays. 

Traps which receive the condensed steam and return it to a boiler 
having considerably higher pressure than that acting on the returns 
are known as return traps. They are made in a great variety of styles. 
The general principle of operation is shown in Fig. 446 and described in 
paragraph 356. 

Non-return Traps. 

Non-return traps, as the name implies, are used where the water of 
condensation is not returned to the boiler but is discharged into any 
receptacle having less than boiler pressure. 



658 



STEAM POWER PLANT ENGINEERING 



CLASSIFICATION OF A FEW WELL-KNOWN STEAM TRAPS. 

Float : 

Bucket 



Steam Traps , 



! McDaniel. 
Cookson. 
!Acme. 
Albany. 

( Morehead. 
f Metal 



Expansion 



Volatile-Fluid 



! Columbia. 
Geipel. 
( Dunham. 
( Heintz. 



Differential J . , 

(Siphon. 



352. Float Traps. — Fig. 437 shows a section through a McDaniel 
improved trap, illustrating the principles of the float type. A hollow 




Fig. 437. McDaniel Float Trap. 

sphere C of seamless copper pivoted at E rises and falls with the change 
of water level in the vessel. The discharge valve M is operated by the 
float. When the trap is empty the float is in its lowest position and 
the discharge valve is closed. Water of condensation flows into the 
trap by gravity through opening D to a certain depth, when the float 
opens the discharge valve and the steam pressure acting on the surface 
of the water forces it through outlet S to tank or atmosphere. After 
the water is discharged the float closes the valve and permits the con- 
densation to collect again. A gauge glass indicates the height of water 
in the chamber. 

Unless float traps are well made and proportioned there is a danger 
of considerable steam leakage through the discharge valve, due to 
unequal expansion of valve and seat and the sticking of moving parts. 



SEPARATORS, TRAPS, DRAINS 



659 



The discharge from a float trap is usually continuous, since the height 
of the float, and consequently the area of the outlet, is proportional to 
the amount of water present. When the trap is working lightly, this 
adjustment is apt to throttle the area and create such a high velocity 
of discharge as to cause a rapid wear of valve and seat. This defect is 
more or less evident in all steam traps discharging continuously. For 
this reason all wearing parts should be accessible and readily replaceable. 
353. Bucket Traps. — Fig. 438 shows a section through an " Im- 
proved Acme" steam trap. The water of condensation enters the cast- 
iron vessel at A, filling the space D between the bucket E and the walls 




Fig. 438. Acme Bucket Trap. 

of the trap. This causes the bucket to float and forces valve V against 
its seat (valve V and its stem being fastened to the bucket as indicated). 
When the water rises above the edges of the bucket it flows into it and 
causes it to sink, thereby withdrawing valve V from its seat. This 
permits the steam pressure acting on the surface of the water in the 
bucket to force the water through the annular space H to discharge 
opening G. When the bucket is emptied it rises and closes valve V 
and another cycle begins. By closing valve R the trap is by-passed 
and the condensation blows directly through passage C to discharge G. 
The discharge from this type of trap is intermittent. 

354. Dump or Bowl Traps. — Fig. 439 shows sections through a 
Bundy bowl trap of the " return" type. The water enters the bowl 



660 



STEAM POWER PLANT ENGINEERING 



through trunnion D and rises until its weight overbalances counter- 
weight E and the bowl sinks to the bottom. As the bowl sinks, arm 
G, which is a part of the bowl, rises and engages the nuts N on valve 
stem H and opens valve /, thus admitting live steam pressure on to the 
surface of the water. The trap then discharges like all others. After 
the water is discharged weight E sinks and raises bowl A, which in turn 
closes valve I, and the cycle begins again. Bowl traps are necessarily 
intermittent in their discharge. 




Fig. 439. A Typical Tilting Trap. 

Fig. 450 shows the application of a bowl trap to a receiver where the 
drips are under a vacuum, and Fig. 451 a similar application to an 
engine receiver where the pressure varies from less than atmospheric 
pressure to a pressure of 40 or 50 pounds. 

355. Expansion Traps. — Expansion traps may be divided into two 
groups : 

(1) Those in which the discharge valve is operated by the relative 
expansion of metals and 

(2) Those in which the action of a volatile fluid is utilized. 
Expansion traps will never freeze, as they are open when cold and all 

the water drains out before the freezing temperature is reached. 

Since traps of this type have little capacity for holding water, 5 to 
10 feet of pipe should be provided between the trap and the pipe to be 
drained in order that the condensation may collect and cool. 

Fig. 440 shows the general appearance of a Columbia expansion trap 
in which the valve is operated by the expansion of metallic tubes. 
Water gravitates to the trap through opening marked "inlet," passes 
through brass pipe O, then downward to the main body of the valves 
and back to outlet valve C. Below pipe and parallel to it is an iron 



SEPARATORS, TRAPS, DRAINS 



661 



rod S, at the end of which is the support or fulcrum of lever R. The 
lower end of this lever is connected to the stem of the valve C, so that 
any movement of the lever is communicated to it. When the trap is 
cold, valve C is open and all water of condensation passes out. The 
moment steam enters the pipe it expands. The amount of expansion 
is multiplied several times by the action of the lever R, so that the 



Inlet? 




Outlet 

Fig. 440. A Typical Expansion Trap. 

movement of the valve is much greater than the expansion of the pipe 0. 
The compensating spring D prevents the brass tube from damaging 
itself by excessive expansion. Lever A permits the trap to be blown 
through by hand. 

Fig. 441 shows a section through a Geipel trap in which the valve 
is operated directly by the expansion of two metallic tubes and the 




Fig. 441. Geipel Expansion Trap. 

movement is not multiplied by levers as with the Columbia. The 
lower or brass pipe constitutes the inlet and is connected to the vessel 
to be drained; the upper or iron pipe is the outlet for discharge. The 
two pipes form the sides of an isosceles triangle, the base F of which is 
rigid, while the apex A is free to move in a direction at right angles to 
the linear expansion of the tubes. When cold, the brass pipe is con- 
tracted and the apex, in which the valve seat is placed, is moved down 
so that the valve is open and the water is discharged. As soon as steam 
enters the brass pipe the latter expands and forces the valve seat against 



662 



STEAM POWER PLANT ENGINEERING 



the valve. The trap may be adjusted for any pressure by means of 
the lock nuts E. When it is desired to blow through, the valve may be 
operated by hand by pressing the lever. 

Fig. 442 shows a section through a Dunham trap. It operates upon 
the expansion principle, utilizing a fluid of a volatile character as its 
motive force. The corrugated bronze disk B is rilled with a volatile 
fluid, and expands and contracts according to the pressure exerted by 




r\ 



Fig. 442. Dunham Expansion Trap. 

the fluid. The water enters at the top, surrounds disk B and passes 
through valve opening D to discharge outlet at E. As soon as steam 
strikes the disk B the volatile fluid flashes into a vapor and causes the 
disk to expand. This expansion forces valve D against its seat and 
the discharge ceases. The valve will remain closed until the condensa- 
tion collects and cools the disk B, which then contracts, opens the valve, 
and condensation enters as before. The adjustment, however, is such 
that the discharge may be made continuous instead of intermittent. 




Fig. 443. Heintz Expansion Trap. 



The Dunham trap is claimed to be the smallest trap of its capacity 
on the market. The 1-inch size, having a capacity for draining 10,000 
lineal feet of 1-inch pipe under 60 pounds pressure, weighs but 5 pounds 
and may be connected to the pipe line as if it were a globe valve. 

Fig. 443 shows an internal view of a Heintz steam trap. This works 
on the principle of the volatile-fluid expansion trap but in a different 



SEPARATORS, TRAPS, DRAINS 



663 



manner from any of those described above. The requisite movement 
is obtained by the elongation and contraction of the extremities of a 
bent metallic tube T filled with a highly volatile fluid. This tube is 
inclosed in a cast-iron box and presses against the point of regulating 
screw P. The other extremity of the tube carries the valve and is free 
to move under the action of the variations of temperature. Spring S 
has no connection with the action of the trap. It is used as a simple 
means of holding one end of the expansion 
tube on its pivot. The trap operates as 
follows: Water enters at I, surrounds the 
tube T and passes through the valve to 
the discharge outlet 0. As soon as steam 
enters the chamber the volatile fluid in the 
tube flashes into a vapor and the pressure 
thus created tends to straighten out the 
tube; this forces the valve against its seat 
and the discharge ceases. As the trap cools 
the tube returns to its normal position and 
the discharge valve is opened, thus per- 
mitting the condensation to drain out. 
The adjustment permits of continuous or 
intermittent discharge and of variable 
pressures. 

356. Differential Traps. — Fig. 444 shows 
a cross section through a Flinn differential 
trap. The column of water X acting on 
diaphragm D closes valve V. The water 
entering pipe E and the action of the spring 
equalize column X and open the valve. 
Describing the action in further detail, the 
water of condensation enters at A, fills 
lower chamber Y, pipeX, and receiving 
chamber C up to the level of the top of 

pipe E. This column of water acting on the under side of the diaphragm 
D forces the valve to its seat against the counter pressure of the spring 
S. Any additional water that enters the trap overflows through pipe 
E, filling chamber F and pipe E to sl point about midway of its height, 
where the effect of the column of water in pipe X is balanced. The 
pressure on each side of the diaphragm is then equal, the short column 
in pipe E, aided by the spring, balancing the pressure of the longer 
column in pipe X. Any further increase in the height of the water in 
pipe E causes a depression of the valve V, which allows water to escape 




V Y 

Fig. 444. Flinn Differential Trap. 



664 



STEAM POWER PLANT ENGINEERING 



Outlet 



until the column has fallen to a level a little below the middle of pipe E, 
when this valve closes again. This action is repeated at intervals 
according to the quantity of water entering the trap. So long as the 
water keeps coming in sufficiently large quantities the valve remains 
wide open. 

Fig. 445 gives a general view of a siphon trap which is much used in 
draining low-pressure systems, as, for example, the separator in an 
exhaust steam heating system. It consists essentially of two legs A 
and B, which may be close together or any distance apart but the 
lengths of which must be sufficiently great to prevent pressure acting 
through pipe I from forcing the water out of B. 
C is a vent pipe extending to the air to prevent 
siphoning; is the discharge for the condensed 
steam. In ordinary operation the leg B is filled with 
water which is constantly overflowing, and A with 
steam and water, the total pressure in both legs being 
equal. The siphon trap is applicable for low pres- 
sure only, as it requires approximately 2.3 feet of 
vertical space E for each pound per square inch 
pressure in the pipe. The maximum allowable head 
is represented by vertical distance N. 

357. Location of Traps. — Wherever possible a trap 
should be located so that the condensation will flow 
into it by gravity. This will insure positive drain- 
age. Sometimes, however, the coils, cylinders, or 
pipes to be drained are located in a pit or trench or 
lie on a basement floor where it is impossible to 
set the trap so as to receive the drains by gravity 
without placing it in an inaccessible position. With 
very low pressures this is often unavoidable, but 
with pressures of five pounds or more the trap may be placed above the 
point to be drained. If a trap is set in an exposed place a drain should 
be provided at the lowest point to free the pipe of water when steam 
is shut ofi\ A dirt catcher or strainer should be placed in the pipe 
leading to the trap to prevent scale, etc., from reaching the valve. All 
pockets and dead ends should be drained, and no condensation should 
be allowed to accumulate. High- and low-pressure drips should be kept 
separate. All tanks should have gauge glasses. 

Fig. 446 shows the application of a float trap for automatically return- 
ing water to the boiler. For this purpose the trap must be placed three 
feet or more above the water line in the boiLer, so that the water may 
gravitate to it. Water is forced into the trap from the returns through 



J Dxa'ii 



Dxa'in 

Fig. 445. Simple 
Siphon Trap. 






SEPARATORS, TRAPS, DRAINS 



665 



pipe A until it reaches a level where the float opens the equalizing valve 
V and permits steam from the boiler to enter the trap, thus equalizing 
the pressures. The water then flows into the boiler by gravity through 
check valve D. At the end of discharge the float closes the equalizing 



Equalizing Valve 




Steam Supply 



o 



F^ 



Return Trap 



21 



Boiler* 



Check 1 „ 

Fig. 446. Return Trap. 

valve and another cycle begins. Check valve C prevents the water 
from being forced back to the return pipe. If the pressure in the re- 
turn pipe A is not sufficient to force the water into the trap, a pump or 
another trap may be used to effect this result. Practically any high- 
pressure trap may be converted into a return trap by proper installation 
and an " equalizing" valve. 




TRAP TRAP 



Fig. 447. 



Drainage System for Jackets and Receivers of Triple-expansion 
Pumping Engines. 



Figs. 447 and 448 show different applications of steam traps to the 
receiver coils and jackets of triple-expansion pumping engines. The 
drawings are self-explanatory. 



666 



STEAM POWER PLANT ENGINEERING 



358. Drips under Vacuum. — Conditions frequently make it neces- 
sary to remove condensation from apparatus working under a vacuum, 
as, for example, a primary heater. 



SEPARATOR 




Cf UNDER 



I 




3 



EXPANSION, 



Fig. 448. Drainage System for Jackets and Receivers of Triple-expansion 
Pumping Engines. 

The simplest method is to pipe the drips to the condenser and per- 
mit the condensation to gravitate to it as in Fig. 449. Where this is 
impracticable, as in an installation with the condenser above the heater, 

a steam trap is usually employed. 
Fig. 450 shows the application 
of a Bundy trap to a vacuum or 
primary heater. A close-fitting 
weighted check valve "FT, set to 
open outwards, prevents intake 
of air through the discharge pipe 
while the trap is filling. Con- 
nection E is made from the vent 
underneath the valve stem V 
back to the heater so as to equal- 
ize the pressures. The operation 




Fig. 449. Gravity Drainage; Vacuum Heater. 



is as follows: Condensation gravitates from the heater through check 
C to the body of the trap, the check W being closed. When the bowl 
is full enough to overcome the weight of the counterbalance, it sinks 
and opens up the live-steam valve V. This admits steam to the trap 
through pipe D, which in turn closes check C and forces the water past 
the weighted check W to the discharge tank. After the water is dis- 
charged the bowl returns to its original position and closes valve V, the 
weight closes check W, the vent check equalizes the pressure in the bowl 
and heater, and condensation gravitates to the trap again. 



i 



1/ 



SEPARATORS, TRAPS, DRAINS 



667 



359. Drips under Alternate Pressure and Vacuum. — Occasionally the 
load on an engine is of such a character that the pressure in the receiver 
alternates from a pressure of 30 or 40 pounds absolute to a vacuum of 
varying degree. Where the periods of vacuum operation are very few 
and of short duration, as in the average installation, no attention is 
paid to the vacuum and the condensation is removed by a trap in the 
ordinary way. If, however, the periods are of sufficient duration and 
frequency, the ordinary method is not applicable and the arrangement 
shown in Fig. 451 may be used. The trap is placed below the receiver 




Fig. 450. Method of Draining Heater under Vacuum. 

as indicated. The delivery pipe is provided with a weighted check or 
resistance valve W set so as to open outwards from the trap, also a 
spring water relief valve R. Another weighted check P is placed in 
the line leading from the vent to the atmosphere, and a plain check C 
in the line leading back into the receiver. This arrangement of valves 
permits the venting of the trap after discharge and effectually excludes 
air from the trap when there is less than atmospheric pressure on the 
receiver. With the relief valve set to open at a pressure in excess of 
the maximum receiver pressure it acts as a "stop" in the pipe and the 
water must enter the trap. When the trap discharges, the live steam 
supplied through the pipe attached to the steam valve forces the water 



668 



STEAM POWER PLANT ENGINEERING 



through the weighted check and relief valves into the sewer or receiving 
tank. When working with a vacuum, the pressures in receiver and 
trap are equalized through the vent connection and the condensation 
flows into the trap by gravity. The operation of discharge is the same 
as in the case of pressure. 




Receiver 



5= 




R<tt 



Floor Line 




Fig. 451. Method of Draining Receivers under Alternate Vacuum and Pressure. 

360. The Steam Loop. — Fig. 452 illustrates the principles of the 
" steam loop" for automatically returning high-pressure drips to the 
boiler. In the figure the loop is returning the condensation from a 
steam separator to a boiler above the level of the separator. The 
apparatus is very simple, consisting of a horizontal and two vertical 
lengths of plain pipe placed as indicated. Pipes R and B may be cov- 
ered but " horizontal" A is left uncovered, as its function is that of a 
condenser. The operation is as follows: Circulation is first started by 
opening stop valve at the bottom of the drop leg until steam escapes. 
The valve is then closed and the steam in the horizontal A condenses 
and gravitates to the drop leg B. On account of the slight reduction 
in pressure in the horizontal a mixture of spray and steam flows from 
the separator chamber to the horizontal, and, condensing, gravitates to 
the drop leg. The column of water in the drop leg rises until its static 
head balances the difference in pressure in the riser R and the horizontal. 
In other words, a decrease in pressure in the horizontal produces similar 
effects on the contents of the riser and drop leg but in a degree in- 



SEPARATORS, TRAPS, DRAINS 



669 



versely proportional to their densities. Any further accumulation causes 
an equal amount to pass from the bottom of the column to the boiler, 
since the pressure in the boiler is then less than that at the bottom of 
the column; that is, the steam pressure on the top of the water column 
plus the hydrostatic head H is greater than the pressure in the boiler. 
Once started the process is continuous and requires no further attention. 




Fig. 452. General Arrangement of the Simple "Steam Loop. 



361. The Holly Loop. — In the application of the steam loop where 
many points requiring drainage are connected to many boilers and 
conditions are more complex, some method other than the simple one 
of radiation may be advisable to secure the necessary lower pressure at 
the top of the loop. Such a method is illustrated in Fig. 453. This 
arrangement differs from the simple loop in that all condensation first 
gravitates to a " Holly" receiver (shown in detail in Fig. 454) before 
passing into the "riser." The receiver is placed below the lowest 
point to be drained and serves as a storage for large or unusual quan- 
tities of water and enables the riser to act at a constant rate independ- 
ent of variable discharge into the receiver. Furthermore, the lower 
pressure in the discharge chamber necessary to secure the lifting of the 
mingled steam and water through the riser, instead of being created by 
condensation as in the simple loop, is produced by a reducing valve B 
discharging into the feed-water heater. The operation of the Holly 
loop is as follows: Circulation is started by opening valve D until 
steam appears. Valve D is then closed and the reducing valve is put 
into commission. Condensation from separators, traps, and pipes 
gravitates to the " receiver," from which it is forced into the " riser" 
in the form of a spray. The spraying effect is produced by a series of 



670 



STEAM POWER PLANT ENGINEERING 







SEPARATORS, TRAPS, DRAINS 



671 



holes drilled in pipe A, Fig. 454. From this receiver the spray and 
moisture rise to the " discharge chamber," on account of the lower pres- 
sure at that point, where 
the steam and entrained 
water are separated, the 
water gravitating to the 
bottom of the chamber and 
thence to the drop leg, and 
the steam discharging 
through the reducing valve 
into the heater. The prin- FlG " 454 ' Holly Receiver * 

ciples of operation are exactly the same as in the simple steam loop. 
362. Returns Tank and Pump. — Low-pressure drips in connection 
with heating systems may be returned to the boiler along with the 
condensation from the heating system by a combined pump and receiver 
as shown in Fig. 455. The height of water in the tank controls the 





«2 



Water 
Relief 




^^^^^??^^^?^?^^?^???^???^^?5^^ 



Prlp Trapped 
toSewer"N s J— n- 

Fig. 455. Returns Tank and Pump. 

operation of the pump through the medium of a float and throttle valve. 
This combination of float and balanced throttle valve is sometimes 
called a "pump governor." In the illustration the pump forces the 



672 



STEAM POWER PLANT ENGINEERING 




DISCHARGE 



returns through a closed heater before delivering them to the boiler, 
though they are oftentimes returned directly. The tank is vented 
to the atmosphere to prevent it from becoming "air bound." The 
cold-water supply or make-up water is sometimes discharged into the 
receiving tank as indicated. With open heaters the cold supply is 
ordinarily controlled by a float within the heater itself. 

363. Office Building Drains. — In the power plants of tall office build- 
ings the public sewers are often above the basement level, and it is 
necessary to remove all liquid wastes mechanically. 

The Shone pneumatic ejector has been found to serve this purpose 

effectually. This apparatus 
is placed in a pit in the 
basement floor into which all 
sewage, drips from engines, 
washings from boilers, and 
ground water gravitate, and 
are [automatically dis- 
charged into the street sewer 
by means of compressed air. 
Fig. 456 gives a sectional 
view of a Shone ejector of 
ordinary construction. It 
consists essentially of r a 
closed vessel .furnished with 
inlet and discharge connec- 
tions fitted with check 
valves, A and B, opening 
in opposite directions with 
regard to the ejector. Two 
cast-iron bells, C and D, are 
linked to each other, in 
Fig. 456. Shone Ejector. reverse positions, the rising 

and falling of which control the supply of compressed air through the 
agency of automatic valve E. 

The bells are shown in their lowest position, the supply of compressed 
air is cut off from the ejector, and the inside of the vessel is open to the 
atmosphere. The sewage gravitating into the ejector raises the bell C, 
which in turn actuates the automatic valve E, thereby closing the con- 
nection between the inside of the ejector and the atmosphere and open- 
ing the connection with the compressed air. The air pressure expels 
the contents through the bell-mouthed opening at the bottom and the 
discharge valve B into the main sewer. Discharge continues until the 




SEPARATORS, TRAPS, DRAINS 



673 



level falls to such a point that the weight of the sewage retained in 
the bell D is sufficient to pull it down, thereby reversing the automatic 
valve. This cuts off the supply of compressed air and reduces the 
pressure to that of the atmosphere. 

The positions of the bells are so adjusted that compressed air is not 
admitted until the ejector is full, and is not allowed to exhaust until 
emptied down to the discharge level; thus the ejector discharges a 
fixed quantity each time it operates. 




Fig. 457. Radiator Drains, Single Gravity System. 

Two ejectors, each of a capacity suitable for handling the average 
flow of tributary sewage and so arranged that they can work either 
independently or together, are usually installed at each ejector station. 

The main sanitary sewer of the building usually discharges directly 
into the ejectors, the surface water, drips, etc., being collected in a 
neighboring sump. The latter is connected to the sanitary sewer 
through a trap or back-water valve. 

364. Radiator Drains. — The condensation from steam heating radia- 
tors is invariably drained back to the boiler. In small heating plants 
with steam pressures of from 1 to 10 pounds gauge pressures, the water 
of condensation is ordinarily allowed to gravitate directly to the boiler 
as shown in Fig. 457. In large plants the steam is often circulated below 
atmospheric pressure, in which case the condensation is withdrawn from 



674 



STEAM POWER PLANT ENGINEERING 



the radiators by mechanical means; see paragraphs 380 and 381. Oc- 
casionally small plants are operated below atmospheric pressure. An 
application of the latter is shown in Fig. 458 and the operation is as 
follows : Steam is generated in the boiler at from 2 to 5 pounds gauge 
and, in flowing through the pipes to the radiators, forces the air 




Fig. 458. Radiator Drains, Dunham " Vacuo-Vapor " System. 

entrainment before it through the radiator trap into return tank A. 
From the latter it is discharged to the atmosphere through automatic 
air valve B. After the air has been expelled the steam comes into con- 
tact with the disk of the radiator trap (see Fig. 442) and the supply is 
cut off thereby preventing further discharge to tank A. Condensation 
collects in the radiator until its temperature is lowered sufficiently to 
cause the trap to open and the flow is again established. As soon as 



SEPARATORS, TRAPS, DRAINS 675 

steam strikes the disk of the radiator valve the flow is cut off and the 
cycle is repeated. The flow of steam to tank A eventually causes the 
water level in the boiler to fall until it reaches the mouth of the equalizing 
tube E. As soon as the end of the tube is uncovered steam flows through 
it into tank A. Steam in tank A immediately closes air valve B and 
check valve F; the pressure in the tank becomes the same as that in 
the boiler and the water gravitates to the boiler through check G. The 
water returning to the boiler raises the water line and seals the equalizing 
tube. The steam in tank A condenses and forms a vacuum of varying 
degree over all the return lines. The system is a sealed one and the 
operation continuous. 



CHAPTER XV. 

PIPING AND PIPE FITTINGS. 

365. General. — The advent of high pressures and superheat is re- 
sponsible for the elimination of many of the older systems of piping, 
the tendency being towards greater uniformity in design, particularly 
in electric central-station work. In isolated stations the conditions of 
operation and installation are so variable that each case presents an 
entirely different problem. In any system of piping the fundamental 
object is to conduct the fluid in the safest and most economical manner. 

The material should be the best obtainable and the system so flexible 
that a break-down in one element will not necessitate the closing down 
of the entire plant. On the other hand, flexibility increases the number 
of parts and, unless first cost is of little importance, tends to weaken 
the system as a whole. It is a safe general proposition to say that the 
best pipe and fittings, irrespective of first cost, will prove the most 
economical in the end, but few owners of power plants are willing to 
take this view. 

366. Drawings. — An assembly drawing of the entire installation 
giving the location of all valves and fittings is necessary in order to 
avoid interference, and particularly where a number of fittings are to 
be close together. Detailed drawings should also be provided of each 
division of the piping to facilitate installation, as, for example, the 
high-pressure steam, the exhaust steam, the feed water, the condensing 
water, the oil, the heating, and the sanitary piping. As a rule, lower 
and more uniform bids will be obtained from an isometric or perspec- 
tive sketch, as in Fig. 459, than from conventional plan and elevation 
drawings, due, no doubt, to the greater ease with which the drawing is 
interpreted. A complete set of specifications for a piping system is given 
in paragraph 479 and illustrates the usual practice along this line. 

367. Materials for Pipes and Fittings. — The following materials are 
used in the construction of pipes for steam, water, and gases. 

Average Tensile Strength 

Low-carbon or mild steel 65,000 lbs. per sq. in. 

Wrought iron 50,000 lbs. per sq. in. 

Cast iron, high grade 20,000 lbs. per sq. in. 

Cast steel • • 50,000 lbs. per sq. in. 

Wrought copper 33,000 lbs. per sq. in. 

Brass 18,000 lbs. per sq. in. 

Special alloys and compounds 15,000-60,000 lbs. per sq. in. 

676 



PIPING AND PIPE FITTINGS 



677 



Mild Steel. — The greater, portion of the piping in the average steam 
power plant is of mild steel, lap or butt welded for high pressures and 
riveted for very low pressures and large diameters. Steel pipe is con- 
siderably cheaper than that manufactured from other material and 
fulfills practically all requirements for general service. 




Fig. 459. A Typical Isometric Pipe Drawing. 



Wrought Iron. — " Wrought-iron" pipe in a commercial sense refers 
to mild-steel pipe and unless stress is laid upon the term " puddled iron" 
mild steel is ordinarily furnished. Puddled-iron pipe is not much in 
evidence in steam power plant work since mild steel is cheaper and 
fulfills all requirements. The claims that wr ought-iron pipe resists 
corrosion to a greater extent than mild-steel pipe have not been sub- 
stantiated in practice. Numerous investigations have been made of 
late which show that mild steel is equal if not superior to wrought iron 
in many ways. 



678 STEAM POWER PLANT ENGINEERING 

Cast-iron Pipes. — Cast iron is little used for high-pressure steam 
piping except occasionally in the construction of headers where a num- 
ber of branches lead into a single pipe, in which case the number of 
joints is greatly reduced and the cost considerably less than for wrought- 
iron or steel pipe with numerous fittings and joints. The chief objec- 
tions to cast iron for high-pressure steam are its weight and lack of 
homogeneity. It is mostly used in connection with water service and 
sanitation. 

Cast-steel Pipe. — Cast-steel headers are sometimes used in power 
plants for highly superheated steam, since the material is not affected 
by temperature variations to the same extent as cast iron. High first 
cost and the difficulty of securing castings free from blowholes have 
prevented its more general use. (See also paragraph 127.) 

Copper Pipes. — Copper steam pipes were in common use for many 
years in marine service on account of their flexibility. To increase the 
bursting strength, pipes above 6 inches in diameter were generally 
wound with a close spiral of copper or composition wire. In recent 
years wrought-iron and steel pipe bends have practically superseded 
copper for flexible connections. As a rule the use of copper pipes should 
be avoided on account of the rapid deterioration of the metal under 
high temperatures and stress variations. The cost is prohibitive for 
most purposes and this alone prevents it from being seriously considered 
in the manufacture of pipe. Copper expansion joints are occasionally 
used in low-pressure work. 

Brass Pipes. — Brass is little used in the construction of pipes on 
account of its high cost. It withstands corrosive action much better 
than iron or steel and is often used in connecting the feed main with the 
boiler drum. Special alloys, nickel steel, "ferrosteel," malleable iron, 
and the like have been used in the manufacture of pipes, and possess 
points of superiority over wrought iron and steel for some purposes, as 
for highly superheated steam, but the cost is prohibitive for average 
steam power plant practice. 

Materials for Fittings. — Elbows, tees, flanges, and similar fittings 
are usually made of cast iron, malleable iron, or pressed steel, though 
cast steel, " ferrosteel," and other steel compounds are used to a limited 
extent. Standard cast-iron fittings are recommended for saturated 
steam and for pressures of 100 pounds per square inch or less, and extra 
heavy cast-iron fittings for higher pressures. Malleable-iron fittings 
are lighter and neater than cast-iron and are extensively used for small 
sizes of steam and gas pipe. Cast or pressed steel is recommended for 
very high pressures and superheat. 



PIPING AND PIPE FITTINGS 679 

368. Size and Strength of Commercial Pipe. — Wrought-iron and 
mild-steel pipes are marketed in standard sizes. Those most commonly 
used in steam power plants are designated as 

1. Merchant or standard pipe. 

2. Full-weight pipe. 

3. Large O.D. pipe. 

4. Extra heavy. 

5. Double extra heavy. 

Table 107 gives the dimensions of standard " full-weight' ' pipe, 
which is specified by the nominal inside diameter up to and including 
12 inches and based on the Briggs' standard. Pipes larger than 12 
inches are designated by the actual outside diameter (O.D.), and are 
made in various weights as determined by the thickness of metal 
specified. Manufacturers specify that " full-weight " pipe may have a 
variation of 5 per cent above or 5 per cent below the nominal or table 
weights, but merchant pipe, which is the standard pipe of commerce, 
such as manufacturers and jobbers usually carry in stock, is almost 
invariably under the nominal weight. It varies somewhat among the 
different mills, but usually lies between 5 and 10 per cent under the 
table weight. The smaller sizes of merchant pipe, | inch to 3 inches, 
are butt-welded and the larger sizes are lap-welded. 

Extra heavy and double extra heavy pipe have the same external 
diameter as the standard, but are of greater thickness and hence the 
internal diameter is smaller. Taking the thickness of the standard 
pipe as 1, that of the extra heavy is approximately 1.4 and of the double 
extra heavy 2.8. 

Wrought-iron and steel pipes are ordinarily designed with factors of 
safety of from 6 to 15, with an average not far from 10. The standard 
hydrostatic tests to which the various pipes are subjected at the mills 
are as follows : 

Hydrostatic Pressure, 
Lbs. per Sq. In. 

Standard, butt-welded, J-3 in 600 to 1,000 

Standard, lap-welded, 3-12 in 500 to 1,000 

Extra heavy, butt-welded, §-3 in 600 to 1,500 

Extra heavy, lap-welded, 1^—12 in 600 to 1,500 

Double extra heavy, butt-welded, J-2J in 600 to 1,500 

Double extra heavy, lap-welded, 1J— 8 in 1,200 to 1,500 

The pressure necessary to burst piping is far above anything likely 
to occur in ordinary practice on account of the thickness of material 
necessary to permit of threading. (See Table 106.) 

Riveted Pipes. — For low pressures and large diameters, pipes are 
constructed of thin sheets of boiler steel with riveted joints, the seams 



680 



STEAM POWER PLANT ENGINEERING 



being either longitudinal and circumferential, or spiral. Such pipes 
are not necessarily limited to large sizes and low pressures, though this 
is the usual practice. 

Pipe fittings are classed as screwed or flanged. 

TABLE 106. 

BURSTING PRESSURE OF " STANDARD " MILD-STEEL PIPE.** 





Nominal 


Actual Bursting 


No. of 
Specimen. 


Nominal 


Actual Bursting 


Specimen. 


Diameter, 
Inches. 


Pressure, 
Lbs. per Sq. In. 


Diameter, 
Inches. 


Pressure, 
Lbs. per Sq. In. 


n 


1 


7800 


J7 


3 


3500 


ta 


1 


7700 


J8 


3 


3500 


13 


1 


7700 
Average 7730 


t9 


3 


3000 
Average 3330 


t4 


2 


4950 


§10 


4 


1800 


t5 


2 


4800 


§11 


4 


1700 


16 


2 


5500 






Average 1750 






Average 5080 


§12 


5 


2500 








§13 


5 


2600 
Average 2550 








§14 


6 


3200 



* Tests made at Armour Institute of Technology. 

Specimens were taken at random from a lot of new pipe; length of test specimens, 5 ft. Specimens 
threaded at both ends and capped. 

t Failed at weld. J Failed in body of pipe. § Failed at threaded end. 

369. Screwed Fittings, Pipe Threads. — For screw connections the 
ends of pipes and fittings are threaded to conform to the Briggs or 
United States standard system, as shown in Fig. 460. The end of the 




Fig. 460. Standard U. S. Pipe Thread. 



pipe is tapered 1 to 32 with the axis, the angle of the thread being 
60 degrees and slightly rounded at top and bottom. The proper length 
of perfect threads is given by the formula 

(0.8 D + 4.8) 



in which 



T = 



T = length in inches. 

D = actual external diameter of the tube, inches. 

n = number of threads per inch. 



(253) 



I 



PIPING AND PIPE FITTINGS 



681 



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682 



STEAM POWER PLANT ENGINEERING 



The imperfect portion of the thread is simply incidental to the pro- 
cess of cutting. The object of the taper is to facilitate " taking hold" 
in making up the joint. Table 107 gives the number of threads per 
inch for various sizes of standard pipe. When properly constructed a 
screwed joint will hold against any pressure consistent with the strength 
of the pipe. For example, the ultimate bursting strength of a " stand- 
ard" 2-inch pipe is about 5000 pounds per square inch, while the 
stripping strength of the joint (with perfect threads) is 225,000 pounds. 

TABLE 108. 

STANDARD BOILER TUBES. 
Table of Standard Dimensions. 



Diameter. 


Standard 
Thickness. 


Transverse 

Areas. 


Area of Surface 

per Foot of 

Tube. 


Nominal Weight per Foot — Lbs. 


"c3 

£ 
2 


a 

CD 


5^ 




"3 
a 

o 


a 
a 

3 


"5 

9 


c3 

C 


"2 8 

§•2 


<3 

III 


0> e3 


Sot 


2 

X m S 


H 


c 


!4PQ 
No. 




a 


a 


H 


s 


»H 


c O 

o 






1 ° 


Ins. 


Ins. 


Ins. 


Sq. In. 


Sq. In. 


Sq. Ft. 


Sq. Ft. 












1 


0.810 


13 


.095 


0.785 


0.515 


.262 


.212 


0.90 


1.04 


1.13 


1.24 


1.35 


H 


1.060 


13 


.095 


1.227 


0.882 


.327 


.277 


1.15 


1.33 


1.45 


1.60 


1.74 


l* 


1.310 


13 


.095 


1.767 


1.348 


.392 


.343 


1.40 


1.62 


1.77 


1.96 


2.14 


if 


1.560 


13 


.095 


2.405 


1.911 


.458 


.408 


1.66 


1.91 


2.09 


2.31 


2.53 


2 


1.810 


13 


.095 


3.142 


2.573 


.523 


.474 


1.91 


2.20 


2.41 


2.67 


2.93 


2\ 


2.060 


13 


.095 


3.976 


3.333 


.589 


.539 


2.16 


2.49 


2.73 


3.03 


3.32 


2h 


2.282 


12 


.109 


4.909 


4.090 


.654 


.597 


2.75 


3.05 


3.39 


3.72 


4.12 


2f 


2.532 


12 


.109 


5.940 


5.035 


.720 


.663 


3.04 


3.37 


3.74 


4.11 


4.56 


3 


2.782 


12 


.109 


7.069 


6.079 


.785 


.728 


3.33 


3.69 


4.10 


4.51 


5.00 


3* 


3.010 


11 


.120 


8.296 


7.116 


.851 


.788 


3.96 


4.46 


4.90 


5.44 


5.90 


3* 


3.260 


11 


.120 


9.621 


8.347 


.916 


.853 


4.28 


4.82 


5.30 


5.88 


6.38 


3f 


3.510 


11 


.120 


11.045 


9.676 


.982 


.919 


4.60 


5.18 


5.69 


6.32 


6.86 


4 


3.732 


10 


.134 


12.566 


10.939 


1.047 


.977 


5.47 


6.09 


6.76 


7.34 


8.23 


4* 


4.232 


10 


.134 


15.904 


14.066 


1.178 


1.108 


6.17 


6.88 


7.64 


8.31 


9.32 


5 


4.704 


9 


.148 


19 . 635 


17.379 


1.309 


1.231 


7.58 


8.52 


9.27 


10.40 


11.23 


6 


5.670 


8 


.165 


28.274 


25.250 


1.571 


1.484 


10.16 


11.19 


12.57 


13.58 


14.65 



The threads, however, are often poorly cut and the parts screwed 
together improperly cleaned and lubricated, thus causing leakage 
between the threads. 

370. Flanged Fittings. — In cast-iron pipes, valves, tees, and other 
fittings the flange is always a part of the casting, but for joining the 
two ends of a steel or wrought-iron pipe the flanges may be fastened to 
the pipe in a number of ways. Fig. 461, A to H, illustrates methods 
most commonly used. In A to C the pipes are screwed into cast-iron 
or forged-steel flanges and the two faces, with metallic or composition 
gasket between, are drawn together by bolts. A illustrates the most 



PIPING AND PIPE FITTINGS 



683 



common and inexpensive of flanged joints, which requires no special 
tools and can be made up at the place of erection. It gives satisfactory 
results for pressures of 100 pounds or less, but for higher pressures 
leakage is apt to take place between the threads. The flanges are 

ma km sJlEEzfll 




m^~--^U ES|™|Aja 




l^a— ~KLB 



SCREWED &. PEENED 




^B 





SHRUNK 

Fig. 461. 



RIVETED 

Types of Pipe Flanges. 



sometimes made with a long thread and a recess which can be calked 
with soft metal. A similar joint is made with the pipe screwed be- 
yond the face of the flange and the two faced off together, either plane 
or as shown in B, which is known as a male and female or hydraulic 
joint. This method forms a very reliable joint, since the ends of the 
pipe bear on the gasket, and the gasket is prevented from being blown 
out. An objection lies in the difficulty of opening the line to remove 
the gasket or replace a fitting. C is a modification known as the 



684 



STEAM POWER PLANT ENGINEERING 



tongued and grooved joint, which uses an extremely narrow gasket. 
Such flanges may be subjected to severe strains when the bolts are 
drawn up, owing to the small area of contact. Corrugated copper or 
steel gaskets are recommended, since soft material is apt to be squeezed 
out. In C the ends of the pipe are peened, which is an improvement 



TABLE 109. 

DIMENSIONS OF CAST-IRON PIPE. 





Standard Thickness and Weight. 






Class A. 




Class B. 






Class C. 


Nominal 


100 Feet Head. 




200 Feet Head. 


300 Feet Head. 


Inside 
Diam- 
eter, 
Inches. 


43 Pounds Pressure. 


86 Pounds Pressure. 


130 Pounds Pressure. 


Thick- 


Weight per 


Thick- 


Weight per 


Thick- 


Weight per 




Inches. 






Inches. 






Inches. 










Foot. 


Length. 




Foot. 


Length. 




Foot. 


Length. 


4 


.42 


20.0 


240 


.45 


21.7 


260 


.48 


23.3 


280 


6 


.44 


30.8 


370 


.48 


33.3 


400 


.51 


35.8 


430 


8 


.46 


42.9 


515 


.51 


47.5 


570 


.56 


52.1 


625 


10 


.50 


57.1 


685 


.57 


63.8 


765 


.62 


70.8 


850 


12 


.54 


72.5 


870 


.62 


82.1 


985 


.68 


91.7 


1,100 


14 


.57 


89.6 


1,075 


.66 


102.5 


1,230 


.74 


116.7 


1,400 


16 


.60 


108.3 


1,300 


.70 


125.0 


1,500 


.80 


143.8 


1,725 


18 


.64 


129.2 


1,550 


.75 


150.0 


1,800 


.87 


175.0 


2,100 


20 


.67 


150.0 


1,800 


.80 


175.0 


2,100 


.92 


208.3 


2,500 


24 


.76 


204.2 


2,450 


.89 


233.3 


2,800 


1.04 


279.2 


3,350 


30 


.88 


291.7 


3,500 


1.03 


333.3 


4,000 


1.20 


400.0 


4,800 


36 


.99 


391.7 


4,700 


1.15 


454.2 


5,450 


1.36 


545.8 


" 6,550 


42 


1.10 


512.5 


6,150 


1.28 


591.7 


7,100 


1.54 


716.7 


8,600 


48 


1.26 


666.7 


8,000 


1.42 


750.0 


9,000 


1.71 


908.3 


10,900 


54 


1.35 


800.0 


9,600 


1.55 


933.3 


11,200 


1.90 


1141.7 


13,700 


60 


1.39 


916.7 


11,000 


1.67 


1104.2 


13,250 


2.00 


1341.7 


16,100 


72 


1.62 


1283.4 


15,400 


1.95 


1545.8 


18,550 


2.39 


1904.2 


22,850 


84 


1.72 


1633.4 


19,600 


2.22 


2104.2 


25,250 

















* Adopted standards of Am. Water W'ks Ass'n. The above weights are per length to lay 12 feet, includ- 
ing standard sockets; proportionate allowance to be made for any variation. All weights are approximate. 

Divisions of Riveted Steel Pipes: Power, March 7, 1911, p. 377. 

over the simple screwed joint. D illustrates a shrunk joint. The 
flanges are bored for a shrink fit and forced over the pipe when at a 
red heat. After cooling the end is beaded over into a recess on the face 
of the flange and a light cut taken from both. H shows a modifica- 
tion in which the hub is riveted to the pipe. E illustrates a joint con- 
structed by rolling the pipe into a corrugation in the flange. The end 
of the pipe is then faced off flush. 






PIPING AND PIPE FITTINGS 685 

One of the best commercial joints is illustrated by F and is known 
as the lap joint. The pipe is expanded as indicated and a light cut is 
then taken from the flared ends to insure a tight joint. The flanges 
are loose and permit of considerable flexibility in shifting them through 
various angles. This is sometimes called the Van Stone joint. 

Pipes with flanges welded on the end as in G have proved the most 
reliable of all and though costly are considered the standard for high- 
pressure and high-temperature work. The faces are ordinarily raised 
sV to T V inch inside the bolt holes and ground to a steam-tight fit, so 
that thick gaskets are unnecessary. 

For moderately high pressures and temperatures any of the joints 
when well made will prove satisfactory. For extremely high pres- 
sures and temperatures the lap or welded joints are preferable. 

The comparative costs of various flanges are given in Table 112. 

Table 110 gives the dimensions of standard and extra-heavy flanges 
and fittings as adopted July 10, 1912, by manufacturers, and Table 
111 the dimensions adopted by "The Societies," Oct. 25, 1911. Since 
neither of these two standards has been universally adopted in this 
country they should be used with caution. 

The following explanatory notes refer to tables 110 and 111. The 
societies' notes appear in Roman type and the manufacturers' variations, 
wherever they occur, follow in italics: 

1. Standard or extra-heavy reducing elbows carry the same dimen- 
sions center to face as the regular elbows of the largest straight size. 

2. Standard or extra-heavy tees, crosses, and laterals, reducing on 
run, carry the same dimensions face to face as the largest straight size. 

3. If flanged fittings for lower working pressures than 125 pounds 
are made, they shall conform in all dimensions, except in thickness of 
shell, to this standard, and shall have the guaranteed working pressure 
cast on each fitting. Flanges for these fittings must be of standard 
dimensions. 

4. Where long-turn fittings are specified, it has reference only to 
elbows, which are made in two center-to-face dimensions, to be known 
as "elbows" and "long-turn" elbows, the latter being used only when 
so specified. 

5. All standard-weight fittings must be guaranteed for 125 pounds 
and extra-heavy fittings for 250 pounds working pressure, and each 
fitting must have some mark cast on it indicating the maker and the 
guaranteed working steam pressure. 

6. All extra-heavy fittings and flanges to have a raised surface T V inch 
high inside of the bolt holes for the gasket. 

Standard-weight fittings and flanges to be plain faced. 



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PIPING AND PIPE FITTINGS 



689 



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690 STEAM POWER PLANT ENGINEERING 

Bolts to be i inch smaller in diameter than the bolt holes. 

Bolt holes should straddle the center lines. 

7. Size of all fittings scheduled indicates the inside diameter of the 
ports. For the outside diameter of pipe use the corresponding size 
of the inside diameter fittings. 

7. Size of all fittings schedules indicates the inside diameter of ports 
except for extra-heavy fittings 14 inches and larger, when the port diameter 
is f inch smaller than nominal size. 

8. The face-to-face dimension of a reducer, either straight or eccen- 
tric, shall be equal to the diameter of the larger flange. 

8. The face-to-face dimension of reducers, either straight or eccentric^ for 
all pressures, shall be the same face to face as given in the table of dimensions. 

9. Square head bolts with hexagonal nuts are recommended. 

10. Twin ells, double-branch ells, side-outlet ells, side-outlet tees and 
fourway tees, whether straight sizes or reducing, carry the same dimen- 
sions center to face and face to face as the regular ells and tees. 

10. Twin elbows, whether straight or reducing, carry same dimensions, 
center to face and face to face, as regular straight-size ells and tees. 

Side-outlet elbows and side-outlet tees, whether straight or reducing sizes, 
carry same dimensions center to face and face to face as regular tees having 
the same reductions. 

11. Bull-head tees or tees increasing on outlet will have the same 
center-to-face and face-to-face dimensions as a straight fitting of the 
size of the outlet. 

12. Up to and including the 4-inch size, center-to-face and face-to- 
face dimensions of reducing fittings will be the same as that of a 
straight fitting of the larger opening. 

12. Tees and crosses 9-inch and down, reducing on the outlet or run 
and outlet, use the same dimensions as straight sizes of the larger port. 

Sizes 10-inch and up, reducing on the outlet, are made in two lengths, 
depending on the size of the outlet as given in the table of dimensions. 

If the outlet is larger than that given in table, use dimensions of straight sizes. 

Laterals 3\-inch and down, reducing on the branch, use the same dimen- 
sions as straight sizes of the larger port. 

Sizes 4-inch and up, reducing on the branch, are made in two lengths, 
depending on the size of the branch as given in the table of dimensions. 

If the outlet is larger than that given in the table, use dimensions of the 
straight sizes. 

Y's are special and are made to suit conditions. 

Double-sweep tees are not made reducing on the run. 

Steel flanges, fittings and valves are recommended for superheated steam. 

13. Pipe sizes 14-inch and over refer to outside diameter. 



PIPING AND PIPE FITTINGS 



691 



TABLE 112. 

COMPARATIVE COST OF VARIOUS PIPE FLANGE FITTINGS, 12-INCH PIPE. 

(Circular from the Crane Company.) 





T3 

S 

<v 
o 
oq 


a 

3 
OQ 


o W 
^ too 

a a 

«3 O 


P 

ex o 
»-3 m 


.3 J2 

O 3 

cj O 




1 


73 
5 


Cast iron 


$ 7.40 

8.70 

9.90 

22.40 

26.40 


$16.00 
18.40 

28^40 
32.40 


$18.00 
20.00 

34.00 
38.00 








$13.00 
16.00 
18.00 
25.00 
30.00 


$21.00 
23.40 


Malleable iron 


$22.00 


$33! 00 

37.00 


$ii! 66 




Cast steel 


33.40 


Weldless steel 


37.40 



Any of the above screwed, shrunk, welded, rolled, or single-riveted flanges can be 
furnished with male or female face at $1 .25 extra. 

The screwed or welded flanges can be furnished with tongued or grooved face at 
$1.25 extra. 

Any of the above screwed, shrunk, or single-riveted flanges can be furnished with 
talking recess at $1.25 extra. 



TABLE 113. 

LOSS OF HEAT FROM BARE STEAM PIPE.* 

Still Air. 



Authority of 
Test. 



Barrus 

Do 

Do 

Hudson Beare 
" 130 lbs.".... 



Jacobus 
Brill.... 



Descriptive Refer- 
ences. 



fcPower, Dec, 1901; 
[Trans. A.S.M.E.,vol. 
Ixxin; Stevens Ind., 
IVol. xix, p. 388. 



Stevens Ind., Vol 

xix, p. 388. 
Stevens Ind., Vol 

XVIII. 

Trans. A.S.M.E., 
Vol. xvi. 



6 


8 

1 « 


a? 

1 
% 


S 


'0 
3- 


3 


f Steam 
ed per 
per Hour. 


£ 


fe *> 


* .' 


£ 


c3 


s si 


g . 





05 -3 

c3 «3 


s s? 


H 6 


a . 


fe a 


■8 -8 £ 

3 3 . 


fj 


3 ^ 




85 ^ 






3 O D* 
O O OQ 


oq 


OQ 


OQ 


OQ 

325 


H 
56 


Q 


Ph 


2 


63.57 


82 


268.6 


0.915 


2 


63.92 


149 


365 


63.3 


302.2 


1.150 


10 


98.33 


149 


365 


73.6 


291.7 


1.085 


3.53f 


8.13 


135 


358 


67 


291 


1.050 


2 


50.66 


128 


354 


80.1 


274.6 


0.994 


2 


7.63 


53 


301 


71.2 


229.6 


0.707 


8 


135.4 


110 


344 


75.5 


269 


0.834 



S3.S 
ft . 

• <4-l 



3.01 
3.25 
3.18 
3.10 
3.13 

2.78 

2.71 



* C. P. Paulding, Stevens Indicator, Vol. xix, p. 388. t Outside diameter. 



692 



STEAM POWER PLANT ENGINEERING 



371. Coverings. — Steam pipes, feed-water pipes, boiler steam drums, 
receivers, separators, etc., should be covered with heat-insulating ma- 
terial to reduce radiation losses to a minimum. For most practical 
purposes the loss of heat from a bare steam pipe or drum may be taken 
as 3 B.t.u. per square foot per hour per degree difference in temperature, 
Table 80. The actual loss depends upon the diameter of the pipe, on 
its position whether vertical or horizontal, the nature of the surface, 
and the velocity of the surrounding air currents. For a detailed analysis 
of these various influences, and interesting information on the trans- 
mission of heat, the reader is referred to Paulding's " Steam in Covered 
and Bare Pipes. " 




Fig. 462. Efficiency of Pipe Coverings. 

By properly applying any good commercial covering, from 75 per 
cent to 90 per cent of the heat loss may be prevented. (See Fig. 462 
and Table 114.) 

Example: Required the saving per annum due to covering a pipe 
10 inches in diameter and 100 feet long; steam pressure 150 pounds; 
average temperature of the air 76 degrees F.; cost of covering applied 
65 cents per running foot; efficiency of covering 85 per cent; cost of coal 
$2.50 per ton; plant to operate 14 hours per day and 300 days per year. 

The temperature of steam at 150 pounds pressure = 366 degrees F. 



PIPING AND PIPE FITTINGS 



693 



Difference of temperature between the steam and air = 366 — 76 
= 290 degrees F. 

Loss per square foot per hour, bare pipe = 3 X 290 = 870 B.t.u. 

Loss per square foot per day, bare pipe = 870 X 14 = 12,180 B.t.u. 

Loss per square foot per year, bare pipe = 12,180 X 300 = 3,654,000 
B.t.u. 

100 lineal feet of 10-inch pipe have an external surface of 282 square 
feet. Therefore the loss per year from the bare pipe is 

2-82 X 3,654,000 = 1,030,000,000 B.t.u. (approx.). 
TABLE 114. 

EXPERIMENTS ON STEAM-PIPE COVERINGS. 

("Condensation of Steam in Covered and Bare Pipes " [Paulding].) 



Kind of Covering. 



Hair felt 

Do 

Remanit for interme- 
diate pressure. 
Remanit for high pres- 
sure. 

Mineral wool 

Champion mineral wool 

Rock wool 

Asbestos sponge felted 

Do 

Do 

Magnesia 

Do 

Do 

Do 

Do 

Do 

Asbestos, Navy Brand 

Do 

Do 

Manville sectional. . . . 

Do 

Do 

Asbestos air cell 

Do 

Asbestos fire felt 

Do 

Do 

Fossil meal. 

Riley cement 



Diam. 
of Test 
Pipe, 
Inches. 



Thick- 
ness of 
Cover- 
ing, 
Inches. 



0.96 
0.82 
0.88 

1.30 

1.30 

1.44 
1.60 
1.125 



.375 
14 
12 
09 
25 
08 
00 
19 
20 
125 
375 
70 
31 
25 



1.12 
0.96 
1.30 
1.00 
0.99 
0.75 
0.75 



Temperatures 
F. 



Steam 



302 

348.3 

304.5 

306.6 

344.1 
346.1 
344.1 
364.8 
364.8 
309.2 
388.0 
354.7 
344.1 
310.9 
365.2 
365.2 
309.2 
365.2 
365.2 
345.5 
354.7 
388.0 
388.0 
303.3 
344.7 
354.7 
307.4 
347.1 
347.9 



71.4 
69.0 
73.3 

76.1 

58.3 
74.3 
63.0 
60.7 
62.8 



79 

72 

80 

66 

81 

64.6 

66.0 

79.4 

64.6 

66.8 

78.3 

80.1 

72.0 

72.0 

72.3 

79.0 

80.1 

72.5 

75.3 

74.3 



B.T.U. per 

Hour per 

Square Foot 

of Pipe 

Surface. 



Total. 



89.6 
117.9 
100.3 

83.7 



81 

86 

72.0 
145.0 

85.0 

59.7 
147.0 
155.8 
106.6 

69.8 
155.0 
103.0 

69.9 
176.0 
112.0 

93.4 
157.0 
143.0 
166.0 
165.5 
133.5 
198 
180.0 
238.0 
260.0 



Per 
Degree 
Differ- 
ence. 



0.387 
0.422 
0.434 

0.363 

0.284 
0.317 
0.256 
0.477 
0.248 
0.260 
0.465 
0.567 
0.384 
0.304 
0.515 
0.347 
0.304 
0.585 
0.375 
0.394 
0.572 
0.453 
0.525 
0.716 
0.502 
0.721 
0.766 
0.876 
0.950 



Date 
of 

Test 



1901 
1894 
1901 

1901 

1894 
1894 
1894 
1901 
1901 
1901 
1896 
1896 
1895 
1901 
1901 
1901 
1901 
1901 
1901 
1894 
1896 
1896 
1896 
1901 
1894 
1896 
1901 
1894 
1894 



Testing Ex- 
pert. 



Jacobus 

Brill 

Jacobus 

Jacobus 

Brill 

Brill 

Brill 

Barrus 

Barrus 

Jacobus 

Norton 

Paulding 

Brill 

Jacobus 

Barrus 

Barrus 

Jacobus 

Barrus 

Barrus 

Brill 

Paulding 

Norton 

Norton 

Jacobus 

Brill 

Paulding 

Jacobus 

Brill 

Brill 



694 STEAM POWER PLANT ENGINEERING 

Assuming a net available heat value of 10,000 B.t.u. per pound for 
the coal, the equivalent coal consumption is 51.5 tons, valued at 51.5 
X $2.50 = $128.75. 

The covering will save 85 per cent of this, or $109.50 per annum. 

The pipe covering applied will cost 100 X $0.65 = $65.00. 

In this case the covering will pay for itself in considerably less than 
a year. 

Pipe covering is applied in sections molded to the required form and 
held to the pipe by bands, or may be applied in a plastic form. The 
former is more readily applied and removed, and is usually adopted 
for pipes, while the valves and fittings are sometimes covered with 
plastic material. Piping should be tested under pressure before being 
covered, since leaks destroy the efficiency and life of the covering. If 
the surrounding atmosphere is moist the covering should be given two 
or three coats of good paint. Coverings are sometimes applied to cold- 
water pipe to prevent sweating in a humid atmosphere. 

Identification of Power House Pi-ping by Colors: Power and Engr., Apr. 26, 1910, 
p. 752. 

372. Expansion. — One of the most difficult problems in the design 
of a piping system is the proper provision for expansion and contraction 
due to change in temperature. If a pipe is immovably fixed at both 
ends and under no strain when cold, and the temperature is increased, 
as by the admission of steam, it is subjected to a compression propor- 
tional to the rise in temperature (within the elastic limit). For example, 
a 6-inch standard extra-heavy wrought-iron pipe 200 feet long at 66 
degrees F., if heated to 366 degrees F. (the temperature corresponding 
to steam at 165 pounds per square inch absolute pressure), will exert 
an axial force of 

P = EA (h - t) fi. (Mechanics of Engng., Church, p. 218.) (254) 

P = force in pounds. 

E = modulus of elasticity, 30,000,000. 

ti = final temperature, degrees F. 
t = initial temperature. 

li = coefficient of expansion, 0.0000075. 

A = sectional area of the pipe material, 8.5 square inches. 

Hence 

P = 30,000,000 X 8.5 (366 - 66) 0.0000075 
= 573,750 pounds. 

Unless well braced throughout its entire length the pipe will buckle 
and become distorted. If free to expand its length would increase. 



PIPING AND PIPE FITTINGS 695 

The temperature of the pipe is always less than that of the steam on 
account of radiation from the outer surface and varies with the effi- 
ciency of the covering. But ignoring radiation the increase in length 
due to temperature increase is 

l=fi(ti-t) L, (255) 

in which 

I = increase in length, inches. 
L = length of pipe, inches. 

Other notations as in (254). 
Substituting in (254), fa = 366. 
t = 66. 

n = 0.0000075. 
L = 2400. 

I = 0.0000075 (366 - 66) 2400 
= 5.4 inches. 
The total increase in length will be the sum of the elongation due to 
pressure and that due to increase in temperature. 

Since the forces produced by expansion are practically irresistible, 
the pipe, is invariably allowed to expand freely by suitable means so 
as not to strain the connections. The coefficients of expansion per 
degree difference in temperature for various pipe materials are given 
in Table 115. 

Headers less than 50 feet in length usually require no special pro- 
visions for expansion, provided the ends are free and the leads to and 




c o 

Fig. 463. Types of Expansion Pipe Bends. 

from the header are not too short, the pipe usually being anchored at 
the middle and permitted to expand in either direction. Free expan- 
sion of the feeders may be provided for 

1. By long radius bends, as in Fig. 463. 

2. By double-swing screwed fittings, as in Fig. 464, or 

3. By packed expansion joints, Fig. 465. 

Where practicable the long radius bends will prove most satis- 
factory. The radius of the bend should not be less than 5 diameters 
of the pipe, and larger if possible. The length of straight pipe at the 
end of each bend should not be less than twice the diameter of the 
pipe measured from the face of the flange. 



696 



STEAM POWER PLANT ENGINEERING 



On account of the great strains to which the joints of pipe bends are 
subjected, the welded joint, G, Fig. 461, is recommended as giving the 
best results. The next best is the lap joint, F, Fig. 461. 

TABLE 115. 

COEFFICIENTS OF LINEAR EXPANSION PIPING MATERIALS. 



Material. 



Wrought iron and mild steel. . . 

Wrought iron 

Cast iron 

Cast steel 

Hardened steel 

Nickel-steel, 36 per cent Nickel 

Copper, cast 

Copper, wrought 

Lead 

Cast brass 

Brass wire and sheets 

Tin cast 

Tin hammered 

Zinc cast 

Zinc hammered 



Temperature 
Range. 



32-212 
32-572 
32-212 
32-212 
32-212 
32-572 
32-212 
32-572 
32-212 
32-212 
32-212 
32-212 
32-212 
32-212 
32-212 



Mean Coeffi- 
cient per De- 
gree F. 



0.00000656 

0.00000895 

.00000618 

,00000600 

00000689 

00000030 

.00000955 

0.00001092 

0.00001580 

0.00001043 

0. 00001075 

0.00001207 

0.00001500 

0.00001633 

0.00001722 



LINEAR EXPANSION OR CONTRACTION OF CAST IRON IN INCHES PER 
100 FEET, — DEGREES F. 



Temperature Difference. 


Expansion. 


Temperature Difference. 


Expansion. 


100 
150 
200 
250 


0.72 
1.1016 
1.5024 
1.9260 


300 
400 
500 
600 
800 


2.376 
3.360 
4.440 
5.616 

7.872 









Multiply by 1.1 for wrought mild steel. 
Multiply by 1.5 for wrought copper. 
Multiply by 1.6 for wrought brass. 

Fig. 463, A, B, C, D, shows applications of pipe bends to straight pipe 
runs. A is the cheapest and most common arrangement for all sizes of 
pipe. B is a modification for limited center-to-center spaces. C shows 
a common method of taking up expansion in straight runs of pipe of 
very large diameters where the space requirements prohibit the use of 
a single U bend. Here the main runs are connected to manifolds which 
in turn are connected by a number of small U bends, the equivalent 
areas of which correspond to that of the large pipes. This makes a 



PIPING AND PIPE FITTINGS 



697 





UU 



more flexible connection than if a single U bend were used. The ar- 
rangement D does away with the elbows required in A, but is not 
applicable to pipes over 8 inches in diameter. 

Figs. 478 and 479 show applications of pipe 
bends to boiler and header connections. 

Fig. 464 shows a double-swing screwed joint in 
which expansion causes the fittings to turn slightly 
and thus relieve the strain. This method is 
usually adopted where long radius bends are not 
practicable on account of lack of space and where 
screwed fittings are used. 

Slip joints, Fig. 465, are now little used except 
with very large pipes and where space prohibits 
long radius bends. When slip joints are employed 

the pipe must 
be securely an- 

J FRONT ELEVATION 

chored to pre- Fig. 464. "Double-swing" 
- Vent the Steam Expansion Joint. 

pressure from forcing the joint apart 

and at the same time permit the pipe 

^.ip 1 "~ in expanding to work freely in the 

Fig. 465. Slip Expansion Joint. stuffing box. Sagging 

of the pipe on either side, which might cause binding in 
the joint, is prevented by suitable supports. 

Expansion in Steam Pipes: Power, July, 1906, p. 426, Jan., 1904, 
p. 30, March, 1904, p. 160, Oct., 1904, p. 609, Dec, 1900; Am. 
Elecn., 10-432; Engr., U. S. } Feb. 1, 1904, p. 125; Eng. News, 
44-194, 47-468, 50-487; Power, June 2, 1908. 

373. Pipe Supports and Anchors. — Pipe lines must be 
supported to guard against excessive deflection and vibra- 
tion. Supports are conveniently classified as (1) hangers, 
(2) wall brackets, and (3) floor stands. 

Fig. 466 illustrates a type of hanger for suspending 
pipes from I beams. The supports being free to swing, 
no provision for expansion is necessary. A properly de- 
signed hanger may be readily removed without disturbing 
the pipe line, and should be adjustable to facilitate 
" lining up." If of rigid construction the lower end should 
be provided with a roller. 

Fig. 467 gives the details of a wall bracket with rolls 
and roll binder. Supports adjacent to long radius bends 
should be provided with roll binders as illustrated to prevent the pipe 




Fig. 466. 

A Typical Pipe 

Hanger. 



698 



STEAM POWER PLANT ENGINEERING 



from springing laterally, but they may otherwise be omitted. The 
rollers are often made adjustable to facilitate lining up. 

Fig. 468 illustrates a typical floor stand. Pipe lines are usually 
securely anchored at suitable points in a manner similar to that illus- 
trated in Fig. 469, the pipe resting on a saddle and being rigidly clamped 
to the bracket by a flat iron band with ends threaded and bolted. 
This limits expansion to one direction and prevents excessive strain on 
the fittings. 



#1 







Fig. 467. A Typical Wall 
Bracket with Binding Roll. 



Fig. 468. A Typ- 
ical Floor Stand. 



Fig. 469. A Typical Pipe 
Anchor. 



Fig. 470 illustrates a method of suspending and counterbalancing 
expansion loops in a main header and Fig. 471 a flexible support for a 
large vertical exhaust header. 

374. General Arrangement of High-pressure Steam Piping. — The gen- 
eral arrangement of piping depends in a great measure upon the space 
available for engines and boilers. 

The engine and boiler room may be placed 

(1) Back to back, Fig. 473. 

(2) End to end, Fig. 472. 

(3) Double decked, Fig. 480. 

The back-to-back arrangement is the most common and, other things 
permitting, is to be preferred on account of the short and direct con- 
nection between engines and boilers and the ease of enlargement. The 
engine and boiler rooms are separated by a wall, and as much of the 
piping as possible is located in the boiler room. 



PIPING AND PIPE FITTINGS 



699 




Fig. 470. Method of Suspending and Counterbalancing Expansion Loops in 

Steam Mains. 




Fig. 471. Spring Support for 30-inch Exhaust Pipe. 



700 



STEAM POWER PLANT ENGINEERING 



The end-to-end arrangement is ordinarily limited to situations where 

the distribution of space pre- 
cludes the back-to-back system. 
The double-decked arrangement 
is frequently used where ground 
space is limited or expensive. 

Engines and boilers are con- 
nected in a variety of ways 
through steam headers as shown 
in the following examples: 

1. Spider system, Fig. 473. 

2. Single header, Fig. 474. 

3. Duplicate headers, Figs. 475 
and 476. 

4. Loop or ring header, Fig. 
477. 

5. The "unit" system, Fig. 
478. 

The spider system is often used 
in small plants. In this arrange- 
ment all branch pipes are brought 
to one central header which is 
made as short as possible. The 
shortness of such a header mini- 
mizes danger from breakdowns, 
and brings all the principal valves 
close together. 

The single-header system is 
perhaps the most common, since 
it embodies simplicity, low first 
cost, and provision for extension. 
The duplicate system is losing 
favor, since experience shows that 
the extra cost of the duplicate 
mains will usually give better 
returns in continuity of operation 
and maintenance if invested in 
high-grade fittings on a single- 
pipe system. A small auxiliary 

header is used in modern plants where double mains are required; see 

Fig. 476. 




PIPING AND PIPE FITTINGS 



701 



The loop header is well adapted for the power plants of tall office build- 
ings, Fig. 481, in which a large number of steam engines, elevator pumps, 
air compressors, and miscellaneous steam-consuming appliances are 
crowded together in a comparatively small space. 

Large modern power plants are, by the latest practice, divided into 
complete and independent units, as in Fig. 478, each prime mover hav- 
ing its own boiler equipment, coal and ash-handling machinery, feed 
pumps, and piping, operated independently of the rest of the plant, 
though provision is made whereby any boiler equipment may provide 
steam for any prime mover. 



Battery Xo.l 




' - : .~~ T 



Battery No. 2 



Battery No.3 



j.2 No.3 No.4 

Fig. 473. " Spider " System. 



The power plant of the Manhattan Elevated Railway Company, 
New York, is practically divided into eight sections each consisting 
of an engine and eight boilers, the boilers being " double decked" 
(Fig. 480). 

The branch pipes from the upper and lower batteries lead into 18- 
inch headers, the steam from each being conducted to a receiver reservoir 
36 inches in diameter and 20 feet long in the engine-room basement 
directly behind each engine, from which the two high-pressure cylinders 
are supplied. Gate valves are used in each boiler branch, one close to 
the boiler and another near the header, and also in the steam pipes 
near the reservoir. The steam headers for each of the eight units are 
connected by a main which equalizes the pressure and allows a deficiency 
in one unit to be made up from the others. 



702 



STEAM POWER PLANT ENGINEERING 




* 




PIPING AND PIPE FITTINGS 



703 




Fig. 476. Typical Auxiliary Header System. 



704 



STEAM POWER PLANT ENGINEERING 







PIPING AND PIPE FITTINGS 



705 




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706 



STEAM POWER PLANT ENGINEERING 




Detail Plan 



(Power) 



Fig. 479. Details of Boiler Steam Piping, Yonkers Power House of the New York 

Central R.R. 



PIPING AND PIPE FITTINGS 



707 



Figs. 478 and 479 show the general arrangement of the steam piping 
at the Yonkers power house of the New York Central. The turbines 
are connected in pairs by 14-inch loops, each turbine taking steam 
from either of two banks of four boilers. The high-pressure steam 
piping is of mild steel with modified reenforced "Van Stone" joints. 
The high-pressure valves are of the split-disk pattern with semi-steel 
bodies. Expansion is taken up by the long sweep bends. 

Plants using superheated steam are ordinarily piped to supply satu- 
rated steam to the auxiliaries as illustrated in Fig. 482. The boiler 
branch E, leading to the main header, normally supplies superheated 
steam to the engines. C is an auxiliary main supplying the air pumps, 
stoker engines, and other auxiliaries with saturated steam from branch 
pipe D. 




Fig. 480. General Arrangement of Steam Piping, Manhattan Elevated Station. 

375. Main Steam Headers. — Until quite recently it was the usual 
practice to make the area of the steam header equivalent to the com- 
bined areas of the feeders, but the function of the header is now regarded 
as that of an equalizer rather than a storage reservoir. In the various 
large power houses recently built in New York City, with ultimate 
capacities of from 100,000 to 250,000 kilowatts, the largest steam headers 
are not over 16 inches in diameter. In some recent designs the pipes 
leading from the header to the engines are two sizes smaller than called 
for by the engine builders. In this case large receiver separators two 
to four times the volume of the high-pressure cylinder are provided 
near the throttle as in Fig. 480. The pipes between receiver and engine 
are full size. The object of the arrangement is to give (1) a constant 
flow of steam, (2) a full supply of steam close to the throttle, and (3) a 
cushion near the engine for absorbing the shock caused by cut-off. 
With moderately superheated steam and boiler pressures from 125 to 
150 pounds a velocity of 8000 feet per minute is allowed in the header 
and as high as 9000 feet per minute between header and receiver. With 
steam turbines velocities as high as 12,000 feet per minute are per- 



708 



STEAM POWER PLANT ENGINEERING 



^ ^ r ~ 1 








"""—• ~" -"A 



PIPING AND PIPE FITTINGS 



709 



missible, provided the pipe is less than 50 feet in length and practically 
free from sharp bends. Main headers are ordinarily constructed of 
mild steel, though cast-iron and cast-steel headers are not uncommon. 




Fig. 482. Overhead Boiler Piping, Quincy Point Power Plant of the Old Colony St. 
Ry. Co., Quincy Point, Mass. 

Cast headers permit of fewer joints and are well adapted to situations 
where a number of branches are closely grouped as in Fig. 473. 

The proper arrangement and number of valves in the main header 
and feeders has been a subject of much consideration. Figs. 472 to 483 




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Fig. 483. Steam Header and Branches, Grand Rapids, Grand Haven and Muskegon 

Ry. Co. Power House. 

show some of the different successful arrangements in recent installa- 
tions. Where two valves are placed in a feeder they should be arranged 
so as not to form a pocket for the accumulation of leakage. In a number 



710 STEAM POWER PLANT ENGINEERING 

of recent installations, Fig. 479, the valve nearest the boiler is of the 
" automatic stop and check" type, its function being the automatic 
cutting off of the steam from the header should the pressure in the boiler 
suddenly drop as in case of blowing out a tube. 

Arrangement of Steam Piping: Power and Engr., Jan. 18, 1910, p. 117, Sept. 29, 
1908, p. 523, Feb. 22, 1908; Engr. U. S., Dec. 1, 1904; Mech. Engr., Nov. 4, 1905; 
Power, Feb. 23, 1909, p. 363; Eng. News, Nov. 26, 1903, p. 487; Elec. Rev., Lond., 
Aug. 11, 1899, p. 251; St. Ry. Rev., Jan., 1900, p. 12; Nov., 1904, p. 869. 

376. Flow of Steam in Pipes.* — The several accepted formulas re- 
lating to the flow of steam in pipes have been based upon a few experi- 
ments limited to pipes of small diameter; hence the application of these 
formulas to larger pipes or to conditions other than those under which 
they were deduced is apt to lead to considerable error. In small plants 
extreme accuracy in determining the proper sizes is not necessary; it is 
better to err in the installation of too large a pipe than one too small. 
In larger stations, however, where the pipes are large and the pressure 
is high, the cost of the piping increases very rapidly with the size. For 
example, the cost of 10-inch high-pressure fittings is from 15 to 20 per 
cent greater than 9-inch fittings, and in large installations this first cost 
item may be of considerable importance. 

The simplest and most commonly used formula is based upon an 
allowable steam velocity of 6000 feet per minute, friction and other 
causes of drop in pressure being disregarded; thus, for a velocity of 
6000 feet per minute, 

(f)», (256) 

in which 

d = diameter of the pipe in inches. 
7 = density of the steam in pounds per cubic foot, and 
W = weight of steam flowing in pounds per minute. 
In determining the diameter of the steam pipe opening for recipro- 
cating engines a much lower velocity than 6000 feet per minute is 
assumed, to allow for the various conditions of operation. Average 
practice gives the constant in equation (256) a value of 0.3 instead of 
0.175 when used in this connection. 

Equation (256) gives satisfactory results for pipes under 100 feet in 
length and between 4 and 8 inches in diameter; for larger diameters 
the velocity could be increased with advantage; for smaller diameters 
or greater lengths friction and condensation would cause considerable 
drop in pressure and some one of the approved formulas in Table 116 
should be used instead. 

* See author's paper, Power, June, 1907, p. 377. 



d = 0.175 ' 



PIPING AND PIPE FITTINGS 



711 



A large drop in pressure means a small pipe and high velocity with 
consequent decrease in condensation, but a point is soon reached 
where the economy in the size of pipe is more than offset by the loss 




1000 2000 3000 4000 5000 6000 SCC0 10000 

Kean Telocity, Feet per Hinule 

Fig. 484. Drop in Pressure for Various Velocities and Pipe Sizes. Initial Pressure 100 

Pounds Gauge, Length of Pipe 100 feet. 

in friction. There seems to be no fixed rule for determining the drop 
most suitable for any given set of conditions. In current practice, the 
drop in pressure between boiler and engine ranges from a fraction of 
one pound to four pounds per square inch per 100 feet of pipe, with an 
average between one and two pounds. 



712 



STEAM POWER PLANT ENGINEERING 



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714 



STEAM POWER PLANT ENGINEERING 




POUNDS OF STEAM 



Fig. 485. 



PIPING AND PIPE FITTINGS 



715 



NOMINAL SIZE OF 




PER HOUR 



Fig. 485. 



716 STEAM POWER PLANT ENGINEERING 

Table 116 gives a few of the best-known formulas for the flow of 
steam, and Table 117 a comparison between them with respect to 
velocity, weight discharged, diameter, and the drop in pressure. 

Formula 11, Table 116, is the most commonly accepted, and the 
curves in Fig. 484 are based upon it, assuming a steam pressure of 100 
pounds absolute and pipe lengths of 100 feet. Within the limit of 
12,000 feet per minute velocity and 10 pounds per square inch drop in 
pressure the curves are sufficiently accurate for all practical purposes, 
but beyond this range the results are purely conjectural and may not 
be accurate, as no recorded experiments have been conducted at these 
high velocities or with pipes of large diameters. 

Though applicable directly to pipes 100 feet long with mean pres- 
sure of 100 pounds per square inch absolute, they may be used for any 
length or pressure. For example, for any length other than 100 feet, 
multiply the drop given in the curves by the required length in feet 
and divide by 100. For any pressure other than 100 pounds abso- 
lute, divide the drop given in the curves by 0.2271 (density of steam 
in pounds) and multiply by the density of steam at the required pres- 
sure. 

Table 118 is the table ordinarily used in connection with the flow of 
steam and is calculated from equation 11. Table 119 is based upon 
equations (4) to (12). The results differ slightly from those in Table 1 18, 
though the latter is more comprehensive. The left-hand half of Table 
119 gives the discharge in pounds per minute for pipes of various 
diameters corresponding to drop of pressure as given on the right-hand 
side in the same horizontal line; e.g., a 6-inch pipe 100 feet long dis- 
charges 371 pounds of steam per minute for a drop of 16.4 pounds at 
100 pounds pressure. 

The curves in Fig. 485 offer a simple means of calculating velocities, 
discharge, and size of pipe for various conditions of flow. The curves 
are plotted for saturated steam only. For superheated or moist steam 
substitute from steam tables pressures corresponding to densities and 
follow chart as directed. 

Example: Allowing a velocity of 5000 feet per minute, what size of 
pipe is necessary to deliver 8000 pounds of steam per hour at 120 
pounds gauge pressure? 

Trace 5000 feet velocity line to 120 pounds gauge line. From inter- 
section follow horizontally to "8000 pounds of steam per hour" and 
read nearest size of pipe, viz., 4 inches. 

Example: Find velocity of steam in a 6-inch pipe delivering 20,000 
pounds of steam per hour at 85 pounds gauge pressure. 

Trace the line representing 20,000 pounds per hour until it inter- 



PIPING AND PIPE FITTINGS 



717 



TABLE 118. 

FLOW OF STEAM THROUGH PIPES (BABCOCK). 





Diameter of Pipe, in Inches. Length of each = 240 diameters. 


Initial Pres- 










sure by 
















Gauge. 


! 


1 


l* 


2 


2* 


3 


4 


Pounds per 
















Square Inch. 












Weight of Steam per Minute, in pounds, with One Pound Loss of Pressure. 


1 


1.16 


2.07 


5.7 


10.27 


15.45 


25.38 


46.85 


10 


1.44 


2.57 


7.1 


12.72 


19.15 


31.45 


58.05 


20 


1.70 


3.02 


8.3 


14.94 


22.49 


36.94 


68.20 


30 


1.91 


3.40 


9.4 


16.84 


25.35 


41.63 


76.84 


40 


2.10 


3.74 


10.3 


18.51 


27.87 


45.77 


84.49 


50 


2.27 


4.04 


11.2 


20.01 


30.13 


49.48 


■ 91.34 


60 


2.43 


4.32 


11.9 


21.38 


32.19 


52.87 


97.60 


70 


2.57 


4.58 


12.6 


22.65 


34.10 


56.00 


103.37 


80 


2.71 


4.82 


13.3 


23.82 


35.87 


58.91 


108.74 


90 


2.83 


5.04 


13.9 


24.92 


37.52 


61.62 


113.74 


100 


2.95 


5.25 


14.5 


25.96 


39.07 


64.18 


118.47 


120 


3.16 


5.63 


15.5 


27.85 


41.93 


68.87 


127.12 


150 


3.45 


6.14 


17.0 


30.37 


45.72 


75.09 


138.61 






Diameter of Pipe, in In 


ches. Length of each= 


240 diameters. 


Initial Pres- 
















sure by Gauge. 


5 


6 


8 


10 


12 


15 


18 


Pounds per 
















Square Inch. 
















Weight 


of Steam per Minute, h 


l Pounds, with One Poi 


md Loss of Pressure. 


1 


77.3 


115.9 


211.4 


341.1 


502.4 


804 


1177 


10 


95.8 


143.6 


262.0 


422.7 


622.5 


996 


1458 


20 


112.6 


168.7 


307.8 


496.5 


731.3 


1170 


1713 


30 


126.9 


190.1 


346.8 


559.5 


824.1 


1318 


1930 


40 


139.5 


209.0 


381.3 


615.3 


906.0 


1450 


2122 


50 


150.8 


226.0 


412.2 


665.0 


979.5 


1567 


2294 


60 


161.1 


241.5 


440.5 


710.6 


1046.7 


1675 


2451 


70 


170.7 


255.8 


466.5 


752.7 


1108.5 


1774 


2596 


80 


179.5 


269.0 


490.7 


791.7 


1166.1 


1866 


2731 


90 


187.8 


281.4 


513.3 


828.1 


1219.8 


1951 


2856 


100 


195.6 


293.1 


534.6 


862.6 


1270.1 


2032 


2975 


120 


209.9 


314.5 


573.7 


925.6 


1363.3 


2181 


3193 


150 


228.8 


343.0 


625.5 


1009.2 


1486.5 


2378 


3481 



For any other length divide 240 by the given length expressed in diameters and multiply 
the tabular quantity by the square root of this quotient, which will give the flow for one pound 
loss of pressure. Conversely, dividing the given length by 240 will give the loss of 
for the flow given in the table. 



718 



STEAM POWER PLANT ENGINEERING 



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PIPING AND PIPE FITTINGS 719 

sects " 6-inch pipe" line. Follow horizontally to "85 pounds gauge" 
and read 7350 feet per minute. 

Example: Allowing a velocity of 6000 feet per minute through an 
8-inch pipe, find the pounds of steam flowing per hour at 100 pounds 
gauge. 

Trace the "6000 feet per minute" velocity line until it intersects 
"100 pounds pressure" line. Follow horizontally to 8-inch pipe line 
and read 32,200 pounds. 

377. Equation of Pipes. — It is frequently desirable to know what 
number of one sized pipes will be equal in capacity to another pipe. 

According to the formulas in Group II, Table 117, the weights dis- 
charged vary with the square root of the fifth power of the diameter; 
that is, the number of pipes equal in capacity to any given pipe may 
be determined from the equation 

N, = di - 4l, . (257) 

in which Ni = number of pipes of diameter di equal in capacity to a 
pipe of diameter d; di and d in inches. 

According to the formulas in Group I, Table 117, the weights dis- 
charged vary as < d 5 -s- ( 1 + -r- ) ( and the equation becomes 

(258) 



(259) 
(260) 



d 




i 6 (d + 3. 

d* Vdj + 3.6 

dS Vd + 3.6 



From (257) and (260) we see that the values of iVi are practically 
the same for either equation when the ratio of d to d\ is small and that 
they differ widely for large ratios. For example, according to (257), 
5.7 eight-inch pipes are equivalent in capacity to one sixteen-inch 
pipe, whereas (259) , gives 6.15. The difference is negligible. Again, 
according to (257), 180 two-inch pipes are equivalent in capacity to 
one sixteen-inch pipe, whereas (260) gives 274. The difference is con- 
siderable. Equation (260) is most commonly accepted and is the basis 
of Table 120. 

378. Friction through Valves and Fittings. — The formulas out- 
lined in Table 116 are strictly applicable only to well-lagged pipes, free 
from bends or obstructions of any kind such as valves or fittings, 
which greatly increase the resistance of the flow of steam. If these 



720 



STEAM POWER PLANT ENGINEERING 



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Q 



PIPING AND PIPE FITTINGS 721 

obstructions must be considered, it is customary to allow for them by 
assuming an added length of straight pipe equivalent in resistance 
to the various fittings and bends. Unfortunately, the few tests which 
have been made for the purpose of determining the resistance of vari- 
ous pipe fittings give discordant results, and in the absence of more 
recent data the rules given by Briggs (" Warming Buildings by Steam") 
are probably as accurate as any. 

According to Briggs, the length of pipe in inches equivalent to the 
resistance of one standard 90-degree elbow is 

L = 7$d~(l+^) (261) 

and to that of one globe valve 

L = lUd~(l+ ^jj. (262) 

The resistance of gate valves is not considered. 

379. Exhaust Piping, Condensing Plants. — The exhaust piping in 
condensing plants is arranged either according to (1) the independent 
or (2) the central condensing system. In the former each engine is 
provided with an independent condenser and air pump. In case the 
vacuum " drops" or it is desired to operate non-condensing, the steam 
is discharged through a branch pipe with relief valve to the atmosphere, 
Figs. 3 and 326. When there are a number of engines in one installa- 
tion the atmospheric pipes lead to a common free exhaust main, which, 
on account of its great size, is ordinarily constructed of light-weight 
riveted steel pipe. The short connection between engine and condenser 
is usually made with lap-welded steel pipe, since riveted joints are apt 
to leak, due to the engine vibrations. In a central condensing plant, 
Fig. 333, the several engines exhaust through a common main into a 
single large condenser. An atmospheric relief valve is usually provided 
in connection with the condenser, and no free exhaust main is necessary. 
Several arrangements of condenser piping are illustrated in Figs. 326 
to 333. 

380. Exhaust Piping, Non-condensing Plant. Webster Vacuum System. 
— In the majority of non-condensing plants the exhaust steam is 
used for heating purposes. One of the best-known systems of exhaust 
steam heating, in which the back pressure on the engine is reduced by 
circulating below atmospheric pressure, is that known as the Webster 
combination system. The general arrangement is illustrated in Fig. 2 
and the principles of operation are described in paragraph 3. It has 
the advantage of affording (1) minimum back pressure on the engine; 
(2) effective and continuous drainage of condensation from supply pipes 



722 



STEAM POWER PLANT ENGINEERING 




PIPING AND PIPE FITTINGS 



723 



and radiators; (3) continuous removal of air and entrained moisture 
from confined spaces; (4) independent regulation of temperature in 
each radiator; (5) continuous return of condensation to the boiler; 
(6) utilization of part of the exhaust steam for preheating the feed 
water; and (7) automatic regulation. Fig. 486 gives a diagrammatic 
arrangement of the piping and appurtenances in a typical installation. 
The characteristic feature of this system is the automatic outlet valve 
attached to each part requiring drainage, which permits both the water 
of condensation and the non-condensable gases to be removed continu- 
ously. The radiator temperature may be regulated by varying the 
quantity of steam supplied, either by hand or automatically by thermo- 
static control. The Webster valve, Fig. 487, enables the vacuum to 
withdraw the water of condensation as fast as it is formed irrespective 



: \\\\\v^W^ 





OUTLET 

Fig. 487. Webster Air Valve. 



Fig. 488. Automatic Vacuum Valve, 
Illinois Engineering Co. 



of the pressure in the radiator; hence the supply may be throttled to 
such an extent that the temperature in the radiator is practically as low 
as that of steam corresponding to the pressure in the vacuum line. 
The small annular space between the inner tube of the float F and the 
guide H permits of a vacuum in the body of the valve. When the water 
from the radiator lifts the float the water is drawn into the returns pipe. 
The valve then returns to its seat and the escape of steam is prevented, 
except such as finds its way through the annular space around the 
guide stem H. An improvement on this valve which prevents the 
escape of steam is illustrated in Fig. 488. When steam is admitted to 
the radiator the condensation flows into the valve, righting the float A 
and sealing the outlet B against the passage of steam; as the valve 
fills with water the buoyancy of the float raises it from its seat and 
permits the water to be drawn out; the float falls and reseats on the 
nipple when about a half-inch of water remains in the valve, thus 
maintaining a water seal. 



724 



STEAM POWER PLANT ENGINEERING 



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1 1 




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Illll^l>l^lllt> 

2 2 2 2 2^.2 sis jjIJ-S *>£ 
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PIPING AND PIPE FITTINGS 



725 



Screen D prevents scale and dirt from entering the valve proper. 
By-pass H is for emergency use in draining off accumulated water in 
the radiator in case the valve becomes stopped up, and permits the 
bonnet to be removed with- suction suction 

out trouble from the accu- 
mulated water. 

381. Exhaust Piping, Non- 
condensing Plants. Paul 
Heating System. — The Paul 
vacuum system differs from 
the Webster in that the 
condensation, and the air 
and non-condensable gases 
are separately handled. 
Referring to Fig. 489, which 
gives a diagrammatic ar- 
rangement of the piping, 
the condensed steam gravi- 
tates to the automatic re- 
turns tank and pump and is 
pumped either directly to 
the boiler or through the 
heater to the boiler. Air 
and vapor are withdrawn from the upper part of the radiator by the 
Paul exhauster or ejector E, and discharged into the returns tank, 
which is vented to the atmosphere for the escape of the non-condensable 

gases. The exhauster receives 




STEAM SUPPLY 

Fig. 490. Paul Exhauster. 



COMPOSITION 




FROM RADIATOR 



its supply of steam through pipe 
0, Fig. 490, which shows the 
general arrangement of this 
apparatus. The piping is in 
duplicate to guard against failure 
to operate. The suction side of 
the exhauster is connected with 
the air pipes A, A, Fig. 489. 
Fig. 491 gives a section through 
the Paul air or vacuum valve 
which prevents steam from blowing into the air pipes and permits only 
air to pass. In Fig. 489 the heating system is piped on what is known 
as the "one-pipe down-feed" principle; i.e., the exhaust steam is first 
conducted to a distributing header in the attic, from which the various 
supply pipes are led to the radiators. The water of condensation 



TO EXHAUSTER 

Fig. 491. Paul Vacuum Valve. 



726 



STEAM POWER PLANT ENGINEERING 



sssss^^ssss 



returns through these same pipes and gravitates to the returns pump. 
Both the supply steam and the condensation flow in the same direction. 
This system is also piped on the " one-pipe up-feed/' the " two-pipe 
up-feed," and the " two-pipe down-feed" principle. The " one-pipe 
up-feed" differs from the system just described in that the steam 
flows upward through the risers and does away with the attic piping. 
The returns, however, flow against the current of steam, and water 
hammer is more likely to occur than with the down-feed system. 
In the two-pipe systems the steam supply pipes or risers conduct 
steam only, and the returns carry the condensation. The one-pipe 
down-feed is cheaper and simpler and practically as efficient as the 
two-pipe system under normal conditions. It is objectionable, however, 
due to the difficulty of draining the radiator with closely throttled 
supply valve, since the velocity of the entering steam prevents the 
water from returning through the same orifice. 

H^ 382. Automatic Temperature Control. — 

Experience shows that a considerable saving 
in fuel may be effected in the heating plants 
of tall office buildings and similar plants by 
automatically controlling the temperature. 
Hand-controlled valves are usually left wide 
open, and when the room becomes too hot 
the temperature is frequently lowered by 
opening the window, resulting in a waste of 
heat which may be considerable in modern 
buildings with hundreds of offices. Many 
successful methods of automatic temperature 
control are available, the usual system 
consisting of thermostats which control the 
supply of heat by means of diaphragm valves, 
the latter taking the place of the usual 
radiator supply valve. 

Fig. 492 shows a Powers thermostat. The 
expansible disk U contains a volatile liquid 
having a boiling point of about 50 degrees 
F. The pressure of the vapor within the disk 
at a temperature of 70 degrees amounts to six 
pounds to the square inch, and varies with 
every change of temperature, causing a variation in the thickness of the 
disk. The disk is attached by a single screw to the lever Q, which rests 
upon the screw F as a fulcrum. The flat spring R holds the lever and 
disk against the movable flange M. Connecting with the chamber N are 




Fig. 492. Section through 
Powers Thermostat. 



PIPING AND PIPE FITTINGS 



727 



AIR SUPPLY 



two air passages H and I. The thermostat is attached by means of two 
screws at the upper end to a wall plate permanently secured to the wall. 
This wall plate has ports registering with H and I, one for supplying 
air under pressure and the other for conducting it to the diaphragm 
motor which operates the valve or damper. Air is admitted through 
H under a pressure of about fifteen pounds per square inch, and its 
passage into chamber N is regulated by the valve J, which is normally 
held to its seat by a coil spring under cap P. K is an elastic diaphragm 
carrying the flange M , with escape valve passage covered by the point 
of valve L. Valve L tends to remain open by reason of the spring. 
When the temperature rises sufficiently 
expansion of the disk U first causes the 
valve to seat, its spring being weaker than 
that above valve J. If the expansive 
motion is continued, valve J is lifted from 
its seat and compressed air flows into 
chamber N, exerting a pressure upon the 
elastic diaphragm K in opposition to the 
expansive force of the disk. If the tem- 
perature falls, the disk contracts and the 
overbalancing air pressure in N results in 
a reverse movement of the flange M, per- 
mitting the escape valve to open and dis- 
charge a portion of the air; thus the air 
pressure is maintained always in direct 
proportion to the expansive power (and 
temperature) of the disk U. The passage 
/ communicates with a diaphragm valve, Fig. 493. The compressed air 
operates the diaphragm against a coiled spring resistance, so that the 
movement is proportional to the air pressure and the supply of steam 
controlled accordingly. The adjusting screw G, squared to receive a 
key, carries an indicator by means of which the thermostat can be set 
to carry any desired temperature within its range, usually from 60 to 80 
degrees. In changing the temperature adjustment lever Q forces the 
disk U closer to or farther from the flange M . 

In connecting up the system compressed air is carried to the thermo- 
stat and diaphragm valves, from a reservoir through small concealed 
pipes. 

In the indirect system of heating the dampers are of the diaphragm 
type and the method of regulation is the same as with the direct system. 

383. Feed-water Piping. — The simplest arrangement of feed-water 
piping may be found in non-condensing plants, in which the feed water 




Fig. 493. 



A Typical Diaphragm 
Valve. 



728 



STEAM POWER PLANT ENGINEERING 



is obtained under a slight head, such as is afforded by the average city 
supply, and is heated in an open heater by the exhaust steam from the 
engine to a temperature varying from 180 to 210 degrees F. The hot 
feed water gravitates from the heater to the pump and then is forced 
to the boiler, or to the economizer if one is used. If a meter is used 



Q 



Q© 



O 




COLO WATER SUPPLY 



Fig. 494. Feed-water Piping; Non-condensing Plant. 

• 

it is generally placed on the discharge side of the pump, and should be 
by-passed to permit it to be cut out for repairs (Fig. 494). Plants 
operating continuously should have feed pumps in duplicate. In some 
cases the returns from the heating system gravitate to the heater and 
only enough cold water is added to make up the loss from leakage, etc. 



tTOP V*LVC 



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L 



n 




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n 






H 






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£ 



' * 1 AUXILIARY rcCO »IM • 



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HZTL 



Fig. 495. Feed-water Piping; Condensing Plant. 

In other cases the returns gravitate to a special " returns tank," from 
which they are pumped directly to the boiler without further heating. 
Occasionally a live-steam purifier is used, especially if the water contains 
a large percentage of calcium sulphate. The feed is then subjected to 
boiler pressure and temperature and the greater part of the impurity 
precipitated before it enters the boiler. Closed heaters are often used 



PIPING AND PIPE FITTINGS 



729 



in place of open heaters. When the supply is not under head a closed 
heater is usually preferred and is placed between the pump discharge 
and the feed main. 

In condensing plants the feed piping is similar to that in non-con- 
densing plants, except that if exhaust steam is used for heating purposes 
it is supplied by the auxiliaries, such as feed pumps, stoker engines, 
condenser engines, and other steam-using appliances. 

In plants having a number of boilers it is customary to run a feed 
main or header the full length of the boiler room and connect it to each 
boiler by a branch pipe. 
This main may be a simple 
header or in duplicate or of 
the " loop "or "ring" type. 
Horizontal tubular boilers 
are frequently arranged in 
one battery with the feed 
main run along the fronts 
of the boilers just above 
the fire doors. Water-tube 
boilers are generally set in 
a battery, and as the ar- 
rangement above would 
block the passageway be- 
tween the batteries, the 
main is run either above 
or under the settings, the 
former being the more 
common. Where a single 
header is used, the feed 
pumps are sometimes 
placed so as to feed into 
opposite ends of the main, 
which is then cut into sections by valves. Another arrangement is to 
place the pumps so as to feed into the middle of the header. With the 
loop arrangement the main is ordinarily cut into sections by valves so 
that the water may be sent either way from the pumps and any defective 
section cut out. With duplicate mains a common arrangement is to 
place one main along the front of the boiler and the other at the rear 
or both overhead as in Fig. 482. Sometimes one main is placed in 
the passageway below the boiler setting and the other on top. 

Standard wrought-iron pipe is usually used for pressures under 100 
pounds and extra heavy pipe for greater pressures. The pipes and 




Fig. 496. Feed-water Piping. 



730 



STEAM POWER PLANT ENGINEERING 



fittings from boiler to main are frequently of brass, and preferably so, 
since brass withstands corrosive action much better than iron or steel. 
Flanged joints should be used in all cases, since the pockets formed by 
the ordinary screwed joints hasten corrosion at those points. (Power, 
June, 1902, p. 4.) 

Fig. 497, A to E, illustrates the various combinations of check valve, 
stop valves, and regulating valve in steam boiler practice. The simplest 




Main Feed Header 

¥iq, 497. Different Arrangements of Valves in Feed-water Branch Pipes. 



arrangement and one sometimes used in plants operating intermittently 
is shown in A. Here there are but two valves between the boiler and 
the main, the check being nearest the boiler and the stop valve at the 
main. The stop valve performs both the function of cutting out the 
boiler and of regulating the water supply. This arrangement is not 
recommended, as any sticking or excessive leaking of the check valve 
will necessitate shutting down the boiler. B shows the most common 
arrangement. Here the check valve is placed between the regulating 
valve and a stop valve as indicated. This permits a disabled check to 
be easily removed while pressure is on the boiler and the main. E shows 
an arrangement whereby both check and regulating valve may be 
removed, and is particularly adapted to boilers operating continuously 
where the regulating valve is subjected to severe usage. In this case 
the stop valves are run wide open and are subjected to no wear. The 
regulating valve most highly recommended is a self-packing brass globe 
valve with regrinding disk. The check valve is ordinarily of the swing 
check pattern with regrinding disk, Fig. 508 (C). Modern practice 
recommends an automatic water relief valve in the discharge pipe im- 
mediately adjacent to each pump to prevent excessive pressure in case 
a valve is accidentally closed in by-passing or in changing over. 

384. Flow of Water through Orifices, Nozzles, and Pipes. — Ber- 
noulli's theorem is the rational basis of most empirical formulas for 
the steady flow of a fluid from an up-stream position n to a down-stream 
position m, thus ("Mechanics of Engineering," Church, p. 706): 



PIPING AND PIPE FITTINGS 731 



P V 2 p yn2 

7 2g y 2g 



all losses of head 
occurring between 
n and m 



(263) 



in which 

V = velocity in feet per second at the point considered. 

P = pressure in pounds per square foot. 

Z = potential head in feet of the fluid. 

7 = density of the fluid, pounds per cubic foot. 

g = acceleration of gravity. 

V 2 
Each loss of head will be of the form K ■=— in which K is the coefficient 

2g 

of resistance to be determined experimentally. The loss of head due to 

skin friction is expressed : 

H = ^ d xf g ,, . (264) 

in which 

/ = the coefficient of friction of the fluid in the pipe. 
I = length of the pipe in feet. 
d = diameter of the pipe in feet. 
Other notations as in (263). 

Discharge from a circular vertical orifice with sharp corners: 

Q = CA V2g~h, (265) 

in which 

Q = cubic feet per second. 

C = coefficient, varying from 0.59 to 0.65 (Merriman, "Treatise on 

Hydraulics/' p. 118). 
A = area of the orifice, square feet. 

h = head of water in feet. 

g = acceleration of gravity = 32.2. 

Discharge from short cylindrical nozzles three diameters in length, with 
rounded entrance ("Mechanics of Engineering/' Church, p. 690): 

Q = 0.815 A V2g~h. (266) 

Discharge from short nozzles with well-rounded corners and conical 
convergent tubes, angle of convergence 13 J degrees (Church, p. 693) : 

Q =0MAV'2g~h. (267) 

Discharge from cylindrical pipe under 500 diameters in length (Church, 
P- 712): 

Q - 6 - 3 V / (i + o^+ ^ (268) 



732 



STEAM POWER PLANT ENGINEERING 



in which 
/ = coefficient of friction. 

Other notations as above. 

/ varies with the nature of the inside surface, the diameter of the 
pipe, and the velocity of flow. 

Discharge through very long cylindrical pipes ("Mechanics of Engineer- 
ing," Church, p. 715): 

Q = Z-15\I~ (269) 

TABLE OF THE COEFFICIENT / FOR FRICTION OF WATER IN CLEAN 

IRON PIPES. 

(Abridged from Fanning.) 



Velocity in 
Ft. per Sec. 


Diam. 
= £ in. 


Diam. 
= 1 in. 


Diam. 
= 2 in. 


Diam. 
= 3 in. 


Diam. 
= 4 in. 


Diam. 
= 6 in. 


Diam. 

= 8 in. 


= .0417 ft. 


= .0834 ft. 


= .1667 ft. 


= .25 ft. 


= .333 ft. 


= .50 ft. 


= .667 ft. 


0.1 


.0150 


.0119 


.00870 


.00800 


.00763 


.00730 


.00704 


0.3 


.0137 


.0113 


850 


784 


750 


720 


693 


0.6 


.0124 


.0104 


822 


767 


732 


702 


677 


1.0 


.0110 


.00950 


790 


743 


712 


684 


659 


1.5 


.00959 


.00868 


.00757 


.00720 


.00693 


.00662 


.00640 


2.0 


.00862 


810 


731 


700 


678 


648 


624 


2.5 


795 


768 


710 


683 


662 


634 


611 


3.0 


.00753 


.00734 


.00692 


.00670 


.00650 


.00623 


.00600 


4.0 


722 


702 


671 


651 


631 


607 


586 


6.0 


689 


670 


640 


622 


605 


582 


562 


8.0 


663 


646 


618 


600 


587 


562 


544 


12.0 


630 


614 


590 


582 


560 


540 


522 


16.0 


.00618 


.00600 


.00581 


.00570 


.00552 


.00530 


.00513 


20.0 


615 


598 


579 


566 


549 


525 


508 


Velocity in 
Ft. per Sec. 


Diam. 


Diam. 


Diam. 


Diam. 


Diam. 


Diam. 


Diam. 


= 10 in. 


= 12 in. 


= 16 in. 


= 20 in. 


= 30 in. 


= 40 in. 


= 60 in. 


= .833 ft. 


= 1.00 ft. 


= 1.333 ft. 


= 1.667 ft. 


= 2.50 ft. 


= 3.333 ft. 


= 5. ft. 


0.1 


.00684 
673 
659 


.00669 
657 
642 


.00623 
614 
603 










0.3 


".00578' 
567 








0.6 


.00504 


"! 00434* 


.06357* 


1.0 


643 


624 


588 


555 


492 


428 


353 


1.5 


.00625 


.00607 


.00572 


.00542 


.00482 


.00421 


.00349 


2.0 


609 


593 


559 


529 


470 


416 


346 


2.5 


596 


581 


548 


518 


460 


410 


342 


3.0 


.00584 


.00570 


.00538 


.00509 


.00452 


.00407 


.00339 


4.0 


568 


553 


524 


498 


441 


400 


333 


6.0 


548 


534 


507 


482 


430 


391 


324 


8.0 


532 


520 


491 


470 


422 


384 


320 


12.0 


512 


500 


478 


457 


412 


377 


.00313 


16.0 


00502 


.00491 


.00470 


.00450 


.00406 


.00370 




20.0 


498 


485 

























PIPING AND PIPE FITTINGS 733 



H = (o.l 



Loss of head due to friction in water pipes* Weisbach's formula is 
as follows: 

0.01716N LV 2 

in which 

H = friction head in feet. 
V = velocity in feet per second. 
'L = length of pipe in feet. 
d = diameter of pipe in inches. 

William Cox (American Machinist, Dec. 28, 1893) gives a simple 
formula which gives almost identical results: 

„_ gv* + 5v-2)L (97U 

H I200d~ (271) 

Notations as in (270). 

Loss of head due to friction of fittings. Formulas (268) to (271) are 
based on the flow of water through clean straight cylindrical pipes. 
Where there are bends, valves, or fittings in the line the flow is decreased 
on account of the additional resistance. 

These frictional losses are conveniently expressed in feet of water, 
thus: 

H = cf g , (272) 

C having the following values: 

Angles. Class of Valve. 

45 degrees. 90 degrees. Gate. Globe. Angle. 

C 0.182 0.98 0.182 1.91 2.94 

Example: Determine the pressure necessary to deliver 200 gallons of 
water per minute through a 4-inch iron pipe line 400 feet long, fitted 
with four right-angle elbows and two globe valves. The water is to 
be discharged into an open tank. 

A flow of 200 gallons per minute gives a velocity of 

200 X 144 

- = 5 feet per second (7.48 = number of gallons per 
7.4o X ou X \-Z.l Z 

cubic foot, and 12.72 = internal area of the pipe, square inches). 

From the preceding table, / = 0.00618 for V = 5. 

From (272), 

25 
Resistance head of 4 elbows = 0.98 X ttt-: X 4 = 1.52 feet. 

64.4 

* See also, Friction Formulas for Commercial Pipe, by Ira N. Evans, Power, 
July 9, 1912, p. 54. 



/ 4 X 5 2 + 5 X 5 - 2\ .._ 
1200X4 40 ° 



734 STEAM POWER PLANT ENGINEERING 

Resistance head of 2 globe valves: 

25 
- 1 - 91 X^n X2 = 1.48 feet. 
04.4 

Resistance head of all fittings: 

1.52 + 1.48 = 3 feet. 

Substitute V = 5, L = 400, and d = 4 in (271). 

-e 

= 10.25 feet, resistance head of the pipe. 

Total resistance head = 10.25 + 3 = 13.25 feet of water, or 5.75 
pounds per square inch. 

Example: How many gallons of water will be discharged per minute 
through above line with initial pressure of 100 pounds per square inch, 
and what will be the pressure at the discharge end? 

Since / depends upon the unknown V, we may put / = 0.006 for a 
first approximation and solve for V; then take a new value of / and 
substitute again, and so on. 

Substitute / = 0.006, d = ~ , h = 100 X 2.3 = 230, and I = 400 in 

(269): 



Q = 3.15l/ - 335 + 230 
• V 0.006 X 400 

= 1.95 cubic feet per second, corresponding to a velocity 

of 22 feet per second. 

From the preceding table, 

./ = 0.00548 (by interpolation) for V = 22 feet per second. 

From (272) the friction of 4 elbows and 2 globe valves is found to 
be 58 feet for V = 22. 

From (164) a resistance head of 58 feet of water for V = 22 is found 
to be equivalent to 136 feet of straight pipe, thus: 



/4 X 22 2 X 5 X 22 - 2\ T 
1200X4 L ' 



-a 

L = 136. 
Substitute/ = 0.0548, I = 400 + 136 = 536 in (162): 



n ■■o,.i/ 0-33 8 X230 
V -^V 0.0058X536 

= 1.74 cubic feet per second, corresponding to a veloc- 
ity of 19.3 feet per second. 

= 780 gallons per minute. 



PIPING AND PIPE FITTINGS 



735 



If greater accuracy is necessary determine / and L f or V = 19.3 and 
proceed as above. 

The total friction head may be determined from (271), thus: 
'4 X 19.3 2 + 5 X 19.3 - 2" 



H 



=( 



536 



1200 X 4 
= 177 feet of water 
= 77 pounds per square inch. 

The pressure at the discharge end will be 

100 — 77 =23 pounds per square inch. 

Average power plant practice gives the following maximum velocities 
of flow in water pipes: 



Size of Pipe in 
Inches. 


Velocity, Feet per 
Minute. 


Size of Pipe in 
Inches. 


Velocity, Feet per 
Minute. 


J. to J 

* to 1J 
\\ to 3 


50 
100 
200 


3 to 6 
Over 6 


250 
300-400 



385. Stop Valves. — The valves used to control and regulate the 
flow of fluids are the most important element in any piping system. 
A good valve should have sufficient weight of metal to prevent distor- 
tion under varying temperature and pressure, or under strains due to 
connection with the piping; the seats should be easily repaired or re- 
newed; there should be no pockets or projections for the accumulation 
of dirt and scale, and the valve stem should permit of easy and efficient 
packing. Stop valves are made in such a variety of designs that a 
brief description will be given of only a few fundamental types. 

Fig. 498 shows a section of an ordinary globe valve, so called because 
of the globular form of the casing. This type of valve is the most 
common in use. Globe valves are designated as (1) inside screw and 
(2) outside screw, according as the screw portion of the stem is inside 
the casting, Fig. 498, or outside, Fig. 499. The top, or bonnet, may be 
screwed into the body of the valve, Fig. 498, or bolted, Fig. 499. The 
smaller sizes, three inches and under, are usually of the screw-top type 
and the larger of the bolt-top type. Valves with outside yoke and screw 
are preferable to others in that they show at a glance whether the 
valve is open or closed, an advantage in changing from one section to 
another. The disks are made in a variety of forms, the material 
depending upon the nature of the fluid to be controlled. Thus, for 
cold water, hard rubber composition gives good results; for hot water 



736 



STEAM POWER PLANT ENGINEERING 



and low-pressure steam, Babbitt metal; for high-pressure steam, copper 
or bronze; and for highly superheated steam, nickel. The valve bodies 
are of brass for sizes under three inches, cast iron for the larger sizes 
and ordinary pressures and temperatures, and cast steel or semi-steel 
for high temperatures and pressures. Globe valves should always be 
set to close against the pressure, otherwise they could not be opened 
if the valves should become detached from the stem. Globe valves 
should never be placed in a horizontal steam return pipe with the stem 
vertical, because the condensation will fill the pipe about half full before 





Fig. 498. A Typical Globe Valve, 
Screw-top, Inside Screw. 



Fig. 



499. A Typical Globe Valve, 
Bolt-top, Outside Screw. 



it can flow through the valve. Globe valves that are open all the time 
are preferably designed with a self-packing spindle, as in Fig. 499, in 
which the top of shoulder C can be drawn tightly against the under 
surface of bonnet S, thus preventing steam from leaking past the screw 
threads while the spindle is being packed. 

Figs. 500 to 503 show different types of gate or straightway valves. 
These valves offer little resistance to the flow of steam or liquid passing 
through them, and are generally used in the best class of work. Fig. 500 
shows a section through a solid-wedge gate valve with outside screw and 
yoke. This form of outside screw and yoke with stem protruding be- 
yond the hand wheel is a perfect indicator to show whether the valve is 
open or shut, as the hand wheel is stationary and the spindle rises in 
direct proportion to the amount the valve is opened. For these reasons 



PIPING AND PIPE FITTINGS 



737 



outside screw valves are preferable for high-pressure work and especially 
for the larger sizes. The seats are made solid, or removable, and of 
various materials for different pressures and temperatures. Fig. 502 
shows a section through a split-wedge gate valve with parallel faces and 
seats. For the sake of illustration this valve is fitted with inside screw. 
In this design the spindle remains stationary so far as any vertical 
movement is concerned, and the gate or plug, being attached to it by 
means of a threaded nut, rises into the bonnet when the spindle is re- 
volved. It is impossible to tell by its appearance whether this form of 




Fig. 500. A Typical Gate Fig. 501. A Typical Gate 
Valve, Solid-wedge, Screw- Valve, Solid-wedge, Bolt-top, 
top, Outside Screw. Inside Screw. 



Fig. 502. A Typical Gate 

Valve, Split-wedge, Bolt-top, 

Inside Screw. 



valve is open or closed. Valves with inside screw are adapted to situa- 
tions where there is considerable dirt and grit, since the screw is inclosed 
and protected, and excessive wear is thus avoided. Gate valves with 
split gates are more flexible than those with solid gates, and hence are 
less likely to leak. Fig. 503 shows the application of the gate system 
to an angle valve. All high-pressure valves above 8 inches in diameter 
should be provided with a small by-pass valve, as the pressure exerted 
against the disk or gate is very great when the valve is closed and the 
force required to move it is considerable. The by-pass valve also 
facilitates " warming up" the section to be cut in and is more readily 
operated than the main valve. 



738 



STEAM POWER PLANT ENGINEERING 



386. Automatic Non-return Valves. — Fig. 504 shows a section 
through an automatic non-return valve as applied to the nozzle of a 
steam boiler. As will be seen from the illustration it practically 
amounts to a large check valve with cushioned disk. The object of 
this device is the equalization of pressure between the different units of 
the battery, the valve remaining closed as long as the individual boiler 
pressure is lower than that of the header. In case a tube blows out the 
valve closes automatically, owing to the reduction of pressure, and 
prevents the header steam from entering the boiler. It acts also as a 





Fig. 503. Ludlow Angle Valve, Gate Pattern. Fig. 504. Anderson Non-return Valve. 

safety stop to prevent steam being turned into a cold boiler while men 
are working inside, because it cannot be opened when there is pressure 
on the header side only. To be successful, such a valve should not open 
until the pressure in the boiler is equal to that in the header; it should 
not stick and become inoperative nor chatter and hammer while per- 
forming its work. Referring to Fig. 504, tail rod E insures alignment 
and hence prevents sticking; steam space C acts as a dashpot to prevent 
hammering of the valve as it rises, and steam space D acts as a cushion 
and prevents hammering at closing. Lip F is made to enter the opening 
in the seat and reduce wire drawing across the seat. Fig. 479 shows the 
installation of a number of non-return valves at the Yonkers power 
house of the New York Central Railway Company. 



PIPING AND PIPE FITTINGS 



739 



387. Emergency Valves and Automatic Stops. — In large power 
plants it is customary to protect the various divisions of the steam 
piping by emergency valves which may be closed by suitable means at 
any reasonable distance from the valve. The simplest form of emer- 
gency stop is a weighted " butterfly" valve, which is to all intents and 
purposes a weighted check, as illustrated in Fig. 508 (D). The weight 
when supported, say by a cord and pulley, holds the valve open; when 



^L 




the cord is cut or released the weight drops and forces 
the valve shut. The cord may lead to any convenient 
and safe distance from the valve. In applying this 
system of control to steam engines the valve is placed 
in the steam pipe just above the throttle and the weight 



^^ 





Fig. 505. Crane 

Emergency Valve, 
Hydraulic. 



Fig. 506. Anderson Triple- 
duty Emergency Valve. 



Fig. 507. Pilot Valve for 
Anderson Triple-duty 
Emergency Valve. 



held up by a lever controlled by the main governor or preferably by a 
separate governor. Should the engine exceed a certain speed, as in case 
of accident to the regular governor, the lever supporting the weight is 
tripped by the emergency governor and the valve is closed automatically. 
For high pressures a rotating plug valve or cock is preferred to the but- 
terfly type, since it is balanced in all positions. Gate and globe valves 
may be converted into emergency valves by having the stems mechani- 
cally operated by electric motors, hydraulic pistons, and the like. Fig. 
505 shows a section through a Crane hydraulically operated emergency 
gate valve. 



740 



STEAM POWER PLANT ENGINEERING 



Fig. 506 shows a partial section through an " Anderson triple-duty" 
emergency valve, and Fig. 507 a section through the pilot valve. A 
steam connection from the main line to the top of a copper diaphragm 
holds the pilot valve closed because of the large area above the dia- 
phragm. A steam pipe connection from underneath the emergency 
piston of the triple-acting valve also leads to the pilot valve. In case 
a break occurs in the main steam line or branches, the pressure is re- 
moved from the top of the pilot valve, causing it to open, thus exhausting 
the pressure from beneath the emergency piston in the triple-acting 
valve. The boiler pressure on top of the emergency piston causes the 
valve to close. Pilot valves may be located at any desirable places, 
thus affording control from different points. 

In the "Locke automatic engine stop system" the stop valve is 
operated by an electric motor which is controlled by contact points 
operated by a speed-limit device. (See Power, August, 1907, p. 471, 
for a detailed description.) 

388. Check Valves. — Fig. 508, A to D, illustrates the different types 
of check valves in most common use. A is a ball check, B a cup or disk 




(A) 



(B) (C) 

Fig. 508. Types of Check Valves. 



(DJ 



check, C a swing check, and D a weighted check. Occasionally the valve 
body is fitted with a valve stem and handle for holding the disk against 
its seat, in which it is designated as a stop check. In A and B the valve 
seat is parallel to the direction of flow and the valve is held in place 
by its own weight and by the pressure of the fluid in case of reverse flow. 
In the swing check the seat is at an angle of about 45 degrees to the 
direction of flow. The latter construction is preferred as it offers less 
resistance to flow and there is less tendency for impurities to lodge on 
the valve seat. By extending the hinge of the swing through the body 
of the valve, a lever and weight may be attached as in D and the check 
will not open except at a pressure corresponding to the resistance of 
the weight. It thus acts as a relief valve and at the same time prevents 
a reversal of flow. Stop checks are usually inserted in boiler feed lines 
close to the boiler, and when locked, act as any ordinary stop valve and 
permit the piping to be dismantled or the regulating valve to be re- 
ground without lowering the pressure on the boiler. Since the wear 






PIPING AND PIPE FITTINGS 



741 



on check valves is excessive and necessitates frequent regrinding they 
are often mounted with regrinding disks, Fig. 508 (C), which may be 
" ground" against the seat without removing the valve from the line. 

389. Blow-off Cocks and Valves. — The requirements of a good blow- 
off valve are that it shall furnish a free passage for scale and sediment, 
that it shall close tightly so as not to leak, and that it shall open easily 
without sticking or cutting. On account of the rather severe service 
to which such valves are subjected, they are made very heavy, with 
renewable wearing parts. 

Fig. 509 gives a sectional view of a Crane ferrosteel valve. The bon- 
net is easily taken off and the disk removed to be refaced or replaced 

rm 

~0 




Fig. 509. Crane Ferrosteel Fig. 510. A Typical Blow-off Fig. 511. Faber Blow-off 
Blow-off Valve. Cock. Valve. 

by a new one. The old disk is repaired by pouring in a hard Babbitt 
metal and facing it off flush. The seats are of brass and oval on top 
to prevent scale lodging between them and the disk, and are so made 
that they may be removed; but it has been found in practice that there 
is not much cutting of the seat, the damage usually being confined to 
the softer Babbitt metal which faces the disk. 

Fig. 511 gives a sectional view of a Faber valve. When the disk, 
which makes a snug fit in the body of the valve, is in the position shown, 
the boiler discharge is practically shut off and any sediment l} T ing on 
the seat is cleaned off by a jet of steam or water. 

Fig. 510 shows a section through a typical blow-off cock of the straight- 
way taper plug pattern with self-locking cam. Plug cocks are often 
used instead of valves on the blow-off piping. 



742 



STEAM POWER PLANT ENGINEERING 













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PIPING AND PIPE FITTINGS 



743 



Current practice recommends the use' of two valves, or rather one 
valve and one cock, in the blow-off line of each boiler. In most of the 
large stations a blow-off valve and a blow-off cock are installed as in- 
dicated in Fig. 512. The number and size of blow-off cocks are usually 
specified by city or state legislation. (For a description of various types 
of blow-off valves, see Power, Dec. 20, 1910, p. 2228.) 

390. Safety Valves. — Fig. 513 shows a section through the simplest 
form of safety valve. The valve is 
held on its seat against the boiler 
pressure by a cast-iron weight as 
indicated. This type has the advan- 
tage of great simplicity, and can be 
least affected by tampering, since it 
requires so much weight that any 
additional amount which would 
seriously overload it can be quickly 
detected. For high pressure and 
large sizes of boiler this class of valve 
is entirely too cumbersome. 

Fig. 514 shows the general details of the common lever safety valve. 
The valve is held against its seat by a loaded lever, thereby enabling 
the use of a much smaller weight than the " dead- weight " type, since 
the resistance is multiplied by the ratio of the long arm of the lever 




Fig. 513. "Dead-weight" Safety Valve. 




Fig. 514. Common Lever Safety Valve. 

to the short one. The proper position of the weight is determined by 
simple proportion. Safety valves of the "dead-weight" or " lever" 
type are little used in modern practice, and their use is prohibited in 
U. S. marine service and in many states. 

Fig. 515 shows a section through a typical pop safety valve in which 
the boiler pressure is resisted by a spring. This type of valve has 
practically supplanted all other forms. The boiler pressure acting upon 
the under side of valve V is resisted by the tension in spring S. As 
soon as the boiler pressure exceeds the resistance of the spring the valve 



744 



STEAM POWER PLANT ENGINEERING 



BLOW OFF LEVER 



lifts from its seat and the steam escapes through opening 0. The static 
pressure of the steam plus the force of its reaction in being deflected 
from the surface A holds the valve open until the pressure in the boiler 
drops about 5 pounds below that at which the valve is lifted. The 
additional area of valve exposed to pressure when the valve lifts causes 

it to open with a sudden motion which 
has given it its name, and it also closes 
suddenly when the pressure has fallen. 
These valves are arranged so that the 
spring tension may be varied without 
taking them apart, and provision is 
made for lifting the seats by means of 
a lever. The seats are of solid nickel 
in the best designs, to minimize corro- 
sion. 

The commercial rating of a safety 
valve is based upon the area exposed to 
pressure when the valve is closed. 

The number and size of safety valves 

for a given boiler are ordinarily specified 

by city or state legislation. 

The logical method for determining the size of safety valves is to 

make the actual opening at discharge sufficient to take care of all steam 

generated at maximum load, thus: 

Let W = maximum weight of steam discharged, pounds per hour. 

A = effective discharge area, square inches. 

P = boiler pressure, pounds per square inch absolute. 

L = lift of valve, inches. 

a = angle of the valve seat with the horizontal. 

K = coefficient determined by experiment. 

D = diameter of valve, inches. 

According to Napier's rule for the discharge of steam through un- 
restricted orifices 




GOILER CONNECTION 



Fig. 



515. Consolidated Pop 
Safety Valve. 



w = 3600 pA =51.4 PA. 



(273) 



Allowing for restriction of orifice 

W = 51AKPA. 



(274) 



Experiments by Philip G. Darling (Trans. A.S.M.E., Vol. 31, 1909, 
p. 123) gave a practically constant value of K = 0.925. Experiments 
conducted by the Consolidated Safety Valve Company gave the same 






PIPING AND PIPE FITTINGS 



745 



average value for K as determined by Darling. Substituting this value 
of K in (274), 

W = 47.5 PA. 



For a flat-seated valve 
Whence 



or 



A = ttDL. 

W = 149 PDL 

W 
D = 0.0067^-- 



For the almost universal 45-degree seated valve 

A = ttDL sine 45 degrees (approx.) 
= 0.707 DL. 
Whence 

A = 105 PDL 
W 



or 



D = 0.0095 



PL 



(275) 
(276) 
(277) 
(278) 

(279) 
(280) 
(281) 



The present rule of the United States Board of Supervising Inspectors 



is 



a = 0.2074 



w 



(282) 



in which * P 

a = area of the safety valve in square inches per square foot of grate 

surface per hour. 
w = pounds of water evaporated per square foot of grate surface per 

hour. 
Formula (282) is derived by allowing a lift of ^V of the nominal valve 
diameter and taking 75 per cent as the added restriction of a 45-degree 
over a flat seat, thus 



a = (o.75tD Xj|); 
which, substituted in Napier's formula, gives a = 0.2074 



w 



The Consolidated Safety Valve Company's circular gives the follow- 
ing rated capacity of its nickel-seat pop safety valves: 

TABLE 121. 

RELIEVING CAPACITIES, CONSOLIDATED POP SAFETY VALVES, STATIONARY TYPE. 

Pounds of Steam per Hour. 



a 


Gauge Pressures (Lbs. per Sq. In. ) 


w > 
> 


20 


40 


60 


80 


100 


120 


140 


160 


180 


200 


220 


240 


260 


280 


300 


n 


600 


950 


1300 


1650 


2000 


2,340 


2,690 


3,030 


3,380 


3,720 


4,070 


4,410 


4,760 


5,110 


5,470 


2 


880 


1390 


1890 


2400 


2900 


3,400 


3,900 


4,410 


4,910 


5,420 


5,920 


6,430 


6,930 


7,430 


7,940 


2\ 


1100 


1730 


2360 


3000 


3620 


4,250 


4,880 


5,500 


6,140 


6,760 


7,400 


8,030 


8,650 


9,300' 9,900 


3 


1430 


2250 


3070 


3890 


4700 


5,530 


6,350 


7,170 


8,000 


8,800 


9,620 


10,400 


11,200 


12,100 


12,900 


3* 


1810 


2830 


3860 


4880 5910 


6,950 


7,960 


9,020 


10,000 


11,100 


12,100 


13,100 14,200 


15,200 


16,300 


4 


2060 


3240 


4410 


5580 6770 


7,950 


9,120 


10,300 


11,500 


12,600 


13,800 


15,0001 16,200 


17,300 


18,500 


4} 


2480 


3900 


5310 


673018150 


9,570 


11,000 


12,400 


13,800 


15,200 


16,700 


18,1001 19,500 


20,900 


22,400 


5 


2940 


4620 


6300 


7970 I 9650 


11,330 


13,000 


14,700 


16,400 


18,100 


19,700 


21,400 23,100 


24,800 


26,500 



746 



STEAM POWER PLANT ENGINEERING 



391. Back-pressure and Atmospheric Relief Valves. — These valves 
are for the purpose of preventing excessive back pressure in exhaust 
pipes. In non-condensing plants such valves are designated as back- 
pressure valves and in condensing plants as atmospheric relief valves. In 
the former the valve is usually adjusted so that a pressure of one to 
five pounds above the atmosphere is necessary to lift it from its seat; 
in the latter the valve lifts at about atmospheric pressure. They are 
practically identical in construction, differing only in minor details. 
A slight leakage in the back-pressure valve is of small consequence, 
but in an atmospheric relief valve it may seriously affect the degree 
of vacuum and throw unnecessary work upon the air pump, hence it 
is customary to " water-seal" the latter. Fig. 516 shows a section 





Fig. 516. Foster Back-pressure Valve. 



INLET 

Fig. 517. Davis Back-pressure Valve. 



through a typical back-pressure valve. The valve proper consists of 
a single disk moving vertically. The valve stem is in the form of a 
piston or dashpot which prevents sudden closing or hammering. The 
pressure holding the valve against its seat is regulated by a spring. 
When the back pressure becomes greater than atmospheric plus that 
added by the spring, the valve raises from its seat and relieves it. 

Fig. 517 shows a section through a Davis back-pressure valve, in 
which the resisting pressure is varied by means of a lever and weight. 

Fig. 486 shows the application of a back-pressure valve to a typical 
heating system. 

Fig. 518 shows a section through a typical atmospheric relief valve. 
Opening B is connected to the exhaust pipe and opening A leads to 
the atmosphere. Under normal conditions of operation atmospheric 
pressure holds valve V against its seat. Water in groove S " water- 
seals" the seat and prevents air from being drawn into the condenser. 






PIPING AND PIPE FITTINGS 



747 



In case the pressure in pipe B becomes greater than atmospheric it 
lifts valve V from its seat and is relieved. Piston P acts as a dashpot 
and prevents the valve from slamming. 

Fig. 519 shows a section through an atmospheric relief valve in which 
the weight of the valve is counterbalanced or even overbalanced by an 
adjustable weight and lever, thereby permitting the valve to open at 
or below atmospheric pressure, as may be desired. 





Fig. 518. Crane Atmospheric Relief Valve. Fig. 519. Acton Atmospheric Relief Valve. 

392. Reducing Valves. — It is often necessary to provide steam at 
different pressures in the same plant, as in the case of a combined power 
and heating plant. To effect this result the reduction in pressure is 
accomplished by passing the steam through a reducing valve, which is 
but an automatically operated throttle valve. There are many different 
forms, the operation of all being based upon the same general principles. 

In the Kieley valve, Fig. 520, the low-pressure steam acts upon the 
top of flexible diaphragm D, and the weighted lever L (which may be 
adjusted to give the desired reduction in pressure) acts upon the other 
side. The movement of the diaphragm causes the balanced valve V 
at the upper end of the spindle to open or close, as may be necessary 
to maintain the desired lower pressure. Inertia weights T and C prevent 
chattering. 

Fig. 521 shows a section through a class G Foster pressure regulator 
or reducing valve. In operation, steam enters at A and passes through 
the main valve port H to the outlet B. Steam at initial pressure passes 
through port-C to chamber P and thence to the top of piston T through 
port L, opening the main valve U. Steam at delivery pressure passes 
through E and raises the diaphragm V against the pressure of spring R, 
allowing spring W to close the auxiliary valve X. The pressure in 



748 



STEAM POWER PLANT ENGINEERING 



chamber J is then equalized by the reduced pressure in ports G and the 
under side of piston X, and thus allows spring Y to close the main valve 
which is then held to its seat by the initial pressure. Any reduction in 
delivery pressure is transmitted to diaphragm V, and permits spring to 
open auxiliary valve X, thereby admitting steam to the top of piston T, 
as previously explained. The delivery pressure is adjusted by screw D; 
thus increasing the tension of spring R increases the discharge pressure ; 
and vice versa. The adjustment once made, the delivery pressure will 




Fig. 520. Kieley Reducing Valve. 




Fig. 521. Foster Pressure Regulator. 



remain constant, regardless of any variable volume of discharge or of 
the initial pressure, so long as the latter is in excess of the delivery 
pressure. W, Fig. 489, shows the application of a reducing valve to an 
exhaust steam heating system. Live steam is led to the valve through 
pipe A. It will be noted that the pipe leading from the valve to the 
heating system is much larger than the high-pressure supply pipe on 
account of the increase in volume of the low-pressure steam. Reducing 
valves should always be by-passed to permit of repairs without shutting 
down the system. Care should be taken in not selecting too large a 
reducing valve, as the valve lift is very small and the larger the valve 
the less will be the lift for a given weight of flow and consequently the 
greater the wire drawing and erosion of the valve seat. 



PIPING AND PIPE FITTINGS 



749 



393. Foot Valves. — Whenever a long column of water is to be moved 
in either suction or delivery pipe it is customary to place a check valve 
near the lower end of the column to prevent the water from backing 
up when the pump reverses or shuts down. The check valve placed at 
the end of the suction pipe is called a foot valve. Any check valve may 
be used as a foot valve, though practice limits the choice to the disk 






Fig. 522. Types of Foot Valves. 



or flap type as illustrated in Fig. 522. To prevent rubbish from destroy- 
ing the action, a strainer or screen is generally incorporated with the 
body of the valve. A, Fig. 522, illustrates a single-flap, B a multi-flap. 
and C a disk valve composed of a nest of small rubber valves. The 
single-flap are usually made in sizes f to 6 inches, the multi-flap 7 to 
16 inches, and the disk valve in all commercial sizes from f to 36 inches. 
For large sizes, 16 to 36 inches, the multi-disk valve is given preference, 
since a number of the disks may be disabled without destroying its 
operation. 

The Use and Abuse of Globe Valves: Power and Engr., Jan., 1909, p. 10. 
Gate Valves in Steam Pipe Lines: Power and Engr., Feb. 16, 1909, p. 320. 
Types of Check Valves and Their Operation: Power and Engr., July 6, 1909, p. 11. 



CHAPTER XVI. 

LUBRICANTS AND LUBRICATION. 

394. General. — The losses due to the friction of the working parts 
of machinery include considerably more than the mere loss of power, 
namely, the depreciation resulting from wear of bearings, guides, and 
other rubbing surfaces, and the expense arising from accidents traceable 
to excessive friction. The power absorbed in overcoming friction varies 
with the type of plant and the character of machinery and is seldom 
less than 5 per cent and often greater than 30 per cent of the total power 
developed. In large central stations these losses approximate 8 per cent 
and in weaving and spinning mills will average as high as 25 per cent. 
.(Trans. A.S.M.E., 6-465.) These figures refer to properly lubricated 
plants operating under normal conditions. The proper selection of 
lubricant is therefore a very important problem, since, besides the cost 
of the lubricant itself, the loss in power and in wear and tear to machinery 
is no small item. A change of lubricant may frequently result in 
marked increase in economy of operation. Table 122 gives an idea of 
the saving effected in power by the proper selection of lubricants in a 
number of mills. (Power, May 12, 1908, p. 752.) The net financial gain 
depends, of course, upon the cost of the oil. As a general rule a 10 per 
cent reduction in friction horse power will more than equal the cost of 
lubricants for one year. The lubricants most commonly met with in 
power plant practice are conveniently classified as oils, greases, and 
solids, and are of animal, mineral, or vegetable origin. 

Reference books: Archbutt and Deeley, Lubrication and Lubricants; Redwood 
Lubricants; W. M. Davis, Friction and Lubrication; Gill, Oil Analysis; Robinson, 
Gas and Petroleum Engines; Thurston, Friction and Lost Work; Gill, Engine 
Room Chemistry. 

395. Vegetable Oils. — Except for certain special purposes and for 
compounding with mineral oils these possess lubricating properties of 
little practical value, since they decompose at comparatively low tem- 
peratures and have a tendency to become thick and gummy. The 
vegetable oils sometimes employed are linseed, cottonseed, rape, and 
castor. 

396. Animal Fats. — Many animal fats have greater lubricating 
power than pure mineral oils of corresponding viscosity but are objec- 

750 



LUBRICANTS AND LUBRICATION 



751 



tionable on account of their unstable chemical composition. They 
decompose easily, especially in the presence of heat, and set free acids 
which attack metals. They are seldom used in the pure state and 
are usually compounded with mineral oils. The animal products used 
in this connection are tallow, neat's-foot oil, lard, sperm, wool grease, 
and fish oil, the first named being the most important. In cylinder 
lubrication, especially in the presence of moisture, the addition of 2 to 

TABLE 122. 



EXAMPLES OF REDUCTION IN FRICTION DUE TO PROPER SELECTION OF 

LUBRICANTS. 


No. of 


Country. 


Plant. 


Mill Oils. 
Test I. 


New Oils. 
Test II. 


Per Cent 
of Trans- 
mission to 
Full Load. 


Po.wer 
Reductions. 


Test. 


Full 
Load, 
I.H.P. 


Trans- 
mission, 
I.H.P. 


Full 
Load, 
I.H.P. 


Trans- 
mission, 
I.H.P. 


Test 
I. 


Test 
II. 


Full ' 
Load, 

Per 
Cent. 


Trans- 
mission, 
Per 
Cent. 


1 
2 A 


America 

America 

America 

America 

England 

England.. 

England 

England 

Ireland 

Scotland 

Scotland 

Germany 

Germany 

Germany 

Germany 

Russia 


Cotton 

Worsted 

Worsted 

Cotton 

Cotton 

Cotton 

Worsted 

Weaving 

Linen 


543.21 

611.60 

702.90 

786.00 

1408.60 

1428.40 

348.10 

495.00 

110.70 

177.70 

325.10 

263.41 

341.36 

341.36 

1135.20 

1238.80 

642.60 

346.60 

364.70 

465.40 

511.37 

6.74r 

137.80 

84.00 


192.70 

'356\66' 
357.90 
111.10 
146.60 
49.90 
61.80 
161.40 
114.03 
118.24 
141.29 
362.60 

230.70" 

"i.77x 

74.90 
31.60 


481.75 
596.30 
648.70 
758.00 
1301.80 
1358.70 
327.50 
453.60 
93.10 
164.60 
293.50 
239.35 
290.53 
299.30 
1034.20 
1069.10 
596.80 
313.60 
336.80 
390.40 
482.43 

5.12i 
116.00 
65.30 


168.90 


35.4 


35.0 


11.31 

2.50 
7.80 
3.56 
7.60 
4.90 
5.90 
8.40 
15.90 
7.40 
9.70 
9.10 
14.90 
12.30 
8.89 
13.70 
7.10 
9.50 
7.70 
16.20 
5.60 
24.00 
15.80 
22.30 


12.35 


B 










3 










4 A 
B 

5 
6 

7 


319.30 
348.90 
99.50 
127.50 
38.60 
56.10 
147.30 
97.11 
95.67 
119.28 
328.10 


25.3 
25.0 
31.9 
29.6 
45 = 
34.7 
49.6 
43.2 
31.7 
41.3 
31.9 


24.5 

25.7 
30.4 
28.1 
41.4 
34.0 
50.2 
40.5 
32.9 
39.8 
31.7 


10.30 
2.50* 
10.40 
13.00 
22.70 


8 A 
B 
9 

10 A 
B 
11 
12 


Woolen 

Woolen 

Cotton 

Worsted 

Worsted 

Jute 

Cotton 

Cotton 

Cotton 


9.20 
8.70 
14.80 
19.10 
15.571 
9.51 


13 


202.20 


35.9 


33.9 


12.40 


14 






15 












16 


England 

Germany 

England 

England 

England 


Paper 










17 










18 
19 
20 


Brass shop 

Iron shop 

Wood shop. . . . 


1.53x 

68.10 
25.40 


26.2 
54.3 
37.6 


29.8 
58.7 
38.8 


13.80 
9.10 
19.60 



Same oil after nine months' use. 
= Electrical units. 



t Not full load of mill. 



t Morning load. 



5 per cent of acidless tallow seems to make the oil adhere better to the 
metal surfaces and increases the lubricating effect, while the proportion 
is so small that ill effects from corrosion or gumming are scarcely 
perceptible. 

397. Mineral Oils. — These are all products of crude petroleum and 
form by far the greater part of all lubricants. They present a wider 
range of lubricating properties than those derived from animal or 
vegetable sources, the thinnest being more fluid than sperm and the 
thickest more viscous than fats and tallows. They are not easily 
oxidized, do not decompose, become rancid, or contain acids. 



STEAM POWER PLANT ENGINEERING 

Crude American petroleum of specific gravity 0.802 may yield the 
following commercial products. ("Gas and Petroleum Engines," 
W. Robinson.) 

TABLE 123. 



Average 
Percentage. 



Speci fie 
Gravity. 



Boiling 

Point, 

Degrees 

F. 



Light Oils. 

C Cymogene 

Petroleum ether < Rhigolene 

r Gasoline 

( C. Naphtha 

Petroleum spirit < B. Naphtha 

( A. Naphtha (benz.) . . . . 

Bumingoils, kerosene. {^ a 7^ e ero8ene ; ;; ;; 

Fuel oils {For making oil gas; fuel 

( Lubricating oils 

Heavy oils < Paraffin wax 

( Residium 



traces 

0.1 

1-1.5 

10 
2-2.5 
2-2.5 

12-20 
40-55 



17.5 
2 

5-10 



0.590 
,625-. 631 
.635-. 658 

,680-. 700 
.717-. 72 
,742-. 745 

,780-. 785 
,800-. 810 



0.85 

885-. 920 
908 at 60 
deg. F. 



32 
64 

85-155 

140-212 
175-245 
212-265 

300-570 
300-680 
and up- 
wards 



480 and 
upwards 



Mineral lubrication oils may be classified as 

(1) Distilled oils, which are produced by distillation from crude 
petroleum and made pale, amber colored, and transparent by treatment 
with acid and alkali. 

(2) Natural oils, which are prepared from crude petroleum, from 
which grit, suspended and tarry impurities have been removed. They 
are dark and opaque and are rich in lubricating properties. 

(3) Reduced oils, or heavy natural oils, from which the lighter hydro- 
carbons have been evaporated and from which the tarry residue has 
been removed by filtration. 

398. Solid Lubricants. — Dry graphite, soapstone, and mica are 
sometimes used as lubricants, though they are usually mixed with 
grease or oils. They cannot easily be squeezed or scraped from between 
the surfaces, and are consequently suitable where very great weights 
have to be carried on small areas and when the speed of rubbing is not 
high. The coefficient of friction of such lubricants is high, and when 
economy of power is essential better results may be secured by the use 
of liberally proportioned rubbing surfaces and liquid lubricants. Under 






LUBRICANTS AND LUBRICATION 



753 



certain conditions of pressure and speed these lubricants will sustain, with- 
out injury to the surfaces, pressures under which no liquid would work. 
Deflocculated graphite suspended in oil or water, and designated com- 
mercially as "oildag" and "aquadag" respectively, is finding favor 
with many engineers. Graphite in this deflocculated condition remains 
suspended indefinitely in water and oil, readily coheres to the journal, 
has great wearing properties, and is easily applied to the wearing surfaces. 
From numerous and long-continued trials it appears that 0.35 per cent 
serves adequately for all purposes. Temperature curves of deflocculated 
graphite in combination with various carrying fluids are given in Fig. 523. 
For further data pertaining to the curves in Fig. 523 and for an extensive 
discussion on the subject of lubrication consult Lubrication and Lubri- 
cating, by C. F. Maberg, Jour. A.S.M.E., Feb. and May, 1910. 




Time in Minutes 

Fig. 523. Tests of Graphite Mixed with Various Lubricants. 

399. Greases. — Under this name may be included the various com- 
pounds which consist of oils and fats thickened with sufficient soap to 
form, at ordinary temperatures, a more or less solid grease. Those 
usually employed are lime, soda, or lead soaps, made with various fats 
and oils. "Engine" greases are thickened with a soap made from 
tallow or lard oil and caustic soda, and often contain neat's-foot oil, 
beeswax, and the like. For exceptionally heavy pressures, graphite, 
soapstone, and mica are sometimes added to the grease. Table 124 
gives an idea of the characteristics of a number of greases. (Prac. 
Engineer, U. S., Apr., 1911, p. 293.) The friction tests were made on 
a small Thurston oil testing machine, 320 r.p.m. and bearing pressure 
of 240 pounds per square inch of projected area. These results are 
purely comparative under the given conditions of rubbing surfaces, 
speed and pressure. For results of these greases tested on a large 
Olsen oil machine consult reference given above. 

Commercial Lubricating Greases: Prac. Engineer, U. S., Apr., 1911, p. 293; 
Tests of Grease Lubrication, Ibid., p. 295; Am. Mach., Aug. 24, 1911, p. 356; Power, 
Nov. 8, 1910, p. 1998. 



754 



STEAM POWER PLANT ENGINEERING 



TABLE 124. 



Type. 


Class. 


Melting 
Point, 
Deg. F. 


Per Cent 
Soap. 


Kind of 
Soap. 


Per Cent 
Free Acid 
as Oleic. 


Average 
Coefficient 
Friction. 


A Mineral 

B Mineral 

C Mineral 


Summer 

Summer 

Winter 

Winter 

Winter 

Winter 

Summer 


167 
178 
165 
163 
142 
125 
120 
41 


38 

20 

23 

16 

19 
1.4 
2.1 



Lime 

Lime 

Lime 

Lime 

Lime 

Potash 

Potash 


Trace 

0.3 

6.1 


Trace 






0.075 
0.054 
0.063 


D Mineral 

E Mineral 

^Tallow No. 3. . . 
G Tallow No. XX 
H Lard oil 


0.057 
0.046 
0.022 
0.029 
0.011 













Type. 


Final Coefficient 

Friction After 

3-Hr. Run. 


Maximum Temper- 
ature of Bearing 
Above that of 
Room, Degs. F. 


Final Temperature 
of Bearing Above 
that of Room at 
End of 3-Hr. Run, 
Degs. F. 


A Mineral 


0.075 

0.050 
0.063 
0.054 
0.046 
0.012 
0.018 
0.010 


70 
70 
76 
69 
58 
38 
45 
13 


68 


B Mineral 


58 


C Mineral 


65 


D Mineral 


58 


E Mineral 


50 


F Tallow No. 3 


18 


G Tallow No. XX 


32 


H Lard oil 


12 







400. Qualifications of Good Lubricants. — A good lubricant should 
possess the following qualities: 

(1) Sufficient "body" to prevent the surfaces from coming into con- 
tact under conditions of maximum pressure. 

(2) Capacity for absorbing and carrying away heat. 

(3) Low coefficient of friction. 

(4) Maximum fluidity consistent with the "body" required. 

(5) Freedom from any tendency to oxidize or gum. 

(6) A high "flash point" or temperature of vaporization and a low 
congealing or "freezing point." 

(7) Freedom from corrosive acids of either metallic or animal origin. 
Lubricating oils are identified by certain tests which are used by 

refiners in grading and classifying the oils and by consumers in buying 
them. These tests usually cover the following: 

(1) Identification of the oil, whether a simple mineral, animal or 
vegetable oil or a mixture. 

(2) Density or gravity. 

(3) Viscosity. 

(4) Flash point. 



LUBRICANTS AND LUBRICATION 



755 



(5) Burning point, fire test. 

(6) Acidity. 

(7) Coefficient of friction. 

(8) Cold test. 

401. Identification of Oil. — The chemical analysis of oils lies in the 
province of the chemist, but some of the characteristics may be readily 
determined by a few simple tests. To detect admixtures of fatty oils 
in mineral oil a small quantity is heated in a test tube for 15 minutes 
with small pieces of either metallic sodium or caustic potash. If fatty 
oil is present, saponification takes place and the soap formed will rise 
to the top as a semi-solid mass and the amount may be estimated. 
Tarry matter may be detected by dissolving a small quantity of oil in 
from 10 to 20 times its bulk of gasoline; the tar and other insoluble 
matter will separate and collect at the bottom. 

For a number of single tests for identifying the various constituents of com- 
pound or adulterated oils consult " Engineering-Room Chemistry," by Augustus H. 
Gill. 

402. Gravity. — The density or specific gravity is conveniently de- 
termined by means of a hydrometer, which, in the oil trade, is graduated 
according to the Baume scale. The relationship between specific gravity 
and degrees Baume at a temperature of 60 degrees F. may be expressed : 

140 

Specific gravity = , ^ ; ■ (283) 

130 + degrees Baume 

Table 125 gives the specific gravity and gravity Baume of a number 

of lubricating oils. 

TABLE 125. 

SPECIFIC GRAVITY AND GRAVITY BAUME OF A NUMBER OF LUBRICANTS. 



Water 

Cylinder oil 

Cylinder oil 

Heavy engine oil. . 
Medium engine oil 
Light engine oil.. . 
Castor machine oil 

Lard oil 

Sperm oil 

Tallow oil 

Cottonseed oil.. . . 

Linseed oil 

Castor oil (pure) . . 

Palm oil 

Rape-seed oil 

Spindle oil 



Specific Grav- 


Gravity 


Flash Test, 


ity. 


Baume\ 


Degrees F. 


1.000 


10 
24.5 




.9090 


575 


.8974 


26 


540 


.9032 


25.5 


411 


.9090 


24 


382 


.8917 


27 


342 


.8919 


27 


324 


.9175 


23 


505 


.8815 


29 


478 


.9080 


24.5 


540 


.9210 


22 


518 


.9299 


19 


505 


.9639 


15 
25 




.9046 


405 . 


.9155 


23 
33 




.8588 


312 



If 
756 STEAM POWER PLANT ENGINEERING 

403. Viscosity. — Viscosity may be defined as the degree of fluidity 
or internal friction of an oil. It is sometimes called the "body." It 
is determined by a viscosimeter. There are a number of different in- 
struments for this purpose but no recognized standard instrument or 
method, so that "viscosity" conveys no meaning unless the name of 
the instrument, the temperature, and the amount of oil tested are given. 
Nearly all instruments are of the orifice type; that is, the viscosity of 
an oil is taken as the number of seconds required for a given amount, 
usually 50 cubic centimeters, to flow through an orifice at a given tem- 
perature. By "specific viscosity" is meant the ratio of the time re- 
quired for the oil to run out to that of an equal quantity of water at 
60 degrees F. The viscosity of engine oils is usually taken at 70 degrees 
F. and of cylinder oils at 212 degrees F. 

404. Flash Point. — The flash point is determined by heating a sam- 
ple of oil in an open or closed cup at the rate of 15 degrees F. per minute 
until a spark will ignite the vapor. The temperature at which this 
occurs is the flash point. So much depends upon the extent of oil 
surface exposed, size of spark, distance spark is held from the oil at the 
time of ignition, and the dimensions of the cup, that there may be con- 
siderable variation in the flash point as obtained by different experi- 
menters. 

405. Burning Point, or Fire Test. — By continuing the application of 
heat and noting the temperature at which the oil takes fire and con- 
tinues to burn, the burning point is obtained. The higher the tem- 
perature under which the oil must work the higher the fire test required, 
so that it will not decompose or volatilize. Too high a fire test gives an 
oil that does not atomize readily enough to reach all parts of the cylinder. 

406. Acidity. — The presence of free acid is determined by shaking up 
equal quantities of oil and water and testing with litmus paper. An- 
other simple test is as follows: A small quantity of oil is placed in a 
test tube with a little cupric oxide (Cu 2 0) and subjected to a gentle 
heat for three or four hours. The reaction with the copper turns the 
solution green if fatty acid is present and blue if vegetable acid is present. 

407. Cold Test. — The "cold test" is the temperature at which the 
oil will just flow. The sample is solidified by means of a freezing mix- 
ture and the temperature noted when it softens sufficiently to flow. 

408. Friction Test. — The coefficient of friction as determined from 
friction-testing machines is useful in obtaining a comparison of oils 
under the test conditions, but gives little information concerning the 
action of the oil under the widely different conditions found in actual 
practice. Table 126 gives the physical properties of a number of 
lubricating oils, with their particular zone of application. 



LUBRICANTS AND LUBRICATION 757 

409. Atmospheric Surface Lubrication. — In a general sense all jour- 
nals, slides, and "atmospheric" surfaces should be lubricated with 
straight mineral oils (as free from paraffin as possible), except when in 
contact with considerable water, in which case it is advisable to add 
20 to 30 per cent of lard oil. Vegetable oils, paraffin oils, and animal 
oils (except lard oil as above stated) are not recommended for general 
engine and dynamo service. The test requirements of a number of 
classes of lubricants are outlined in Table 89 and represent current 
practice. Bearings, guides, and all external rubbing surfaces may be 
lubricated in a number of ways. (1) They may be given an inter- 
mittent application of oil, as, for example, with an oil can; (2) they may 
be equipped with oil cups with restricted rates of feed; and (3) they 
may he flooded with oil. The relative lubricating values of the systems 
have been estimated approximately as follows (Power, December, 1905) 
p. 750) : 





Coefficient of Fric- 
tion. 


Comparative 
Value. 


Intermittent 


0.01 and greater 
0.01 to 0.012 
0.00109 


72 and less 


Restricted feed 


79 to 86 


Flooded bearing 


100 



410. Intermittent Feed. — Intermittent applications are ordinarily 
limited to small journals, pins, and guides which are subject to light 
pressures and which do not easily permit of oil or grease cups, as, for 
example, parts of the valve gear of a Corliss engine, governors, and link 
work. On account of the labor attached and the frequent doubt 
about the oil reaching the wearing surfaces this method of lubrication 
is limited as much as possible even in the smallest plants. 

411. Restricted Feed. — In the average power plant the major part 
of the lubrication is effected by means of oil cups which are filled at 
intervals by hand or by mechanical means, the oil being fed from the 
cup by dropsy according to the requirements. 

412. Oil Bath. — In large power plants the principal journals and 
wearing parts are supplied with a continuous flow of oil which com- 
pletely " floods" the rubbing surfaces. The oil is forced to the various 
parts either by gravity from an elevated tank or by pressure from a 
pump. After the oil leaves the bearings it flows into collecting pans, 
thence into a receiving and filtering tank, and finally is pumped back 
into an elevated reservoir and used over and over again. The little 
lost -by leakage and depreciation is replenished by the addition of new 
oil to the system. 



758 



STEAM POWER PLANT ENGINEERING 



TABLE 126. 
PHYSICAL CHARACTERISTICS OF A NUMBER OF LUBRICANTS. 

(Power, December, 1905, p. 750.) 



Kind of Oil. 


Use and Adaptation. 


£ $ 

'> bo 


r. 0) 

8 a 


18 . 

CD W 

s 

X5 Eo 


1 s 


Viscosity 
at 70 De- 
grees. 


High-pressure cylinder 
oil. 


For steam cylinders using dry 
steam at pressures from 110 
to 210 pounds. 


25 

to 

24.5 


30 


600 
to 
610 


645 
to 
660 


175 

to 

205 


General cylinder oil . . 


For steam cylinders using dry 
steam at 75 to 100 pounds. 
For air compressor cylinders 
when made from steam-re- 
fined mineral stock and when 
viscosity is 200. 


26 

to 

25.5 


30 


550 
to 
585 


600 
to 
630 


180 
to 
190 


Wet cylinder oil. 
(Remark 1.) 


For use where the steam is moist, 
especially in compound and 
triple expansion engines. 


25.8 

to 
25.3 


30 


560 
to 
585 


600 
to 
630 


150 
to 

185 


Gas engine cylinder oil. 
(Remark 2.) 


For gas engine cylinders. Neu- 
tral mineral oil compounded 
with an insoluble soap to give 
body. 


26.5 


30 


320 


350 


300 


Automobile gas engine 
oil. (Remark 3.) 


For automobile gas engines and 
similar work. 


29.5 


30 


430 


485 


195 


Heavy engine and 
machinery oils. 


For heavy slides and bearings, 
shafting, and horizontal sur- 
faces. 


30.5 

to 
29.5 


30 


400 


440 
to 
450 


170 

to 
195 


General engine and 
machine oils. 


For high-speed dynamos and 
machines. 


30.8 
to 
30 


30 


400 
to 
420 


450 
to 
470 


175 
to 
190 


Fine and light machine 
oils. 


For fine work, from printing 
presses to sewing machines 
and typewriter oils. With a 
cold test of 25° to 28° and a 
viscosity of 140° this makes 
an excellent spindle oil. 


32.5 

to 
30.2 


30 


400 


440 


110 
to 
160 


Cutting and heat dis- 
sipating oils. 
(Remark 4.) 


For cutting tools, screw cutting 
and similar work. 


27 
to 
23 


30 


410 
to 

420 


475 
to 

480 


210 
to 
175 


Refrigerating oils. . 


For ice machinery 


30.2 





200 


225 


165 








Wet service and marine 
oils. (Remark 4.) 


For marine service, or where a 
great deal of moisture must 
be handled. 


28 


30 


430 


475 


230 


Greases 


They are used in special work 
and for heavy pressures mov- 
ing at slow velocities. 

















Remark 1. — May contain not over 2 to 6 per cent of refined acidless tallow oil in the high- 
pressure oils and not over 6 to 12 per cent in the low-pressure oils. 

Remark 2. — The reason for using an insoluble soap such as oleate of aluminum is that it 
is impossible to decompose the soap with a high heat ; the soap, although not a lubricant, is a 
vehicle for carrying some oil. 

Remark 3. — Owing to a lack of body, this oil will not interfere with the sparking by depos- 
iting carbon on the platinum point. 

Remark 4. — May contain 30 to 40 per cent of pure strained lard oil. 



LUBRICANTS AND LUBRICATION 



759 



413. Oil Cups. — Fig. 524 illustrates the application of sight-feed oil 
cups to the crosshead and slides of a reciprocating engine. The oil is 
fed into the cups by hand and 
gravitates to the rubbing surfaces, 
the rate of flow being regulated by 
a needle valve. Cups A and B feed 
directly to the crosshead guides, 
but the oil from cup D flows to 
the bottom orifice 0, from which 
it is wiped by a metallic wick S, 
and carried by gravity to the wrist 
pin. 

414. Telescopic Oiler. — Fig. 525 
shows the application of a telescopic 
oiler to a crosshead and guides. 
and C are sight-feed oil cups, the 
former feeding directly to the top 
guide through the tube S. The oil 
from C flows by gravity through the swing joint into the telescopic tubes 
P, R, and thence to the pin through the lower swing joint as indicated. 




Fig. 524. Oil-cup Lubrication, 
Hand-filled. 





Fig. 525. Nugent's Telescopic Oiler. 



As the crosshead moves back and forth, the pipe P slides into and out of 
pipe R, the oil being thus conducted directly to the pin without wasting. 
A device of this type installed on a high-speed automatic engine at the 



760 



STEAM POWER PLANT ENGINEERING 



Armour Institute of Technology has been in operation for three years 
without cost for repair or renewal. 

415. Ring Oiler. — Small high-speed engines are often oiled by the 
oil-ring system, as illustrated in Fig. 526. The shaft is encircled by 




IttitU/lBWHZZZ, 







Fig. 526. Oil-ring Lubrication. 



O O Ol 



OOO 




Fig. 527. Centrifugal Oiler. 



several loose rings which dip into a bath of oil in the base of the pedestal 
or frame and, rolling on the shaft as it turns, carry oil to the top of the 
shaft where it spreads to the bearings. In some cases the rings are 
replaced by loops of chain. 

Ring Lubrication: Zeit. d. Ver. Deutscher Ing., Aug. 10, 1907. 




Fig. 528. Pendulum Oiler. 



416. Centrifugal Oiler. — Fig. 527 illustrates the application of a 
centrifugal oiler to a side-crank engine. The oil supply is regulated by 
the sight-feed cup C and flows by gravity to the pipe P in line with 



LUBRICANTS AND LUBRICATION 761 

the center of the crank shaft. Centrifugal force throws the oil outward 
through pipe B to the center of the pin D, which is drilled longitudinally 
and radially so as to distribute the oil upon the bearing surface. 

417. Pendulum Oiler. — Fig. 528 illustrates the application of a 
pendulum oiler to the crank pin of a center-crank engine. Oil cups 
and pendulum P are fastened to the crank shaft S by trunnion T. The 
pendulum holds the cup vertical, since the friction of the trunnion is not 
sufficient to revolve it. Oil flows along the center of the crank shaft 
under the head of oil in cup and is thrown outward to bearing B by 
centrifugal force. 

418. " Splash Oiling." — In some high-speed engines the crank, con- 
necting rod, and crossheads are inclosed by a casing, the bottom of 
which is filled with oil to such a depth that at each revolution of the 
crank, the end of the connecting rod is partly submerged. The result 
is that the oil is splashed into every part of the chamber, and the crank 
pin, crosshead pin, and crosshead slides practically run in an oil bath. 

419. Gravity Oil Feed. — Fig. 529 illustrates a simple gravity oil-feed 
system. The oil to the engine is supplied from the oil tank by pipe D 
under pressure corresponding to the height of the tank above the oil 
cups. After performing its function the oil gravitates to the filter and 
from the latter to the oil reservoir, from which it is pumped back to 
the supply tank, the overflow being returned to the reservoir through 
pipe N. Operation is interrupted only when new oil is to be added to 
the system from the barrel through the flexible filling pipe. In case 
the oil tank is put out of commission, or the supply pipe becomes clogged, 
full pump pressure may be used by closing valves R and S and opening 
valve E. The make-up oil is small in amount compared to the quantity 
circulated. The reclaiming and purifying of the oil are essential if the 
bearings are to be flooded, otherwise the cost of oil would be prohibitive. 
At the power house of the South Side Elevated Railway the daily 
circulation (24 hours) of engine oil is approximately 1500 gallons. The 
make-up oil amounts to eight gallons. 

An objection sometimes made to the above system is that the varying 
heights of oil in the supply tank may cause considerable variation in 
pressure at the oil cups, causing them to feed faster when the tank is 
full and slower when the tank is nearly empty. This applies only to 
installations where the supply tank is filled intermittently. 

420. Low-pressure Gravity Feed. — Fig. 530 shows the application 
of a low-pressure oiling system in which the level in the sight feeds is 
kept constant. A is the main supply tank, B 1 and B 2 the upper and 
lower gauges indicating the oil level, C the supply pipe running to the 
engines, and D & small standpipe closed at one end and vented near the 



762 



STEAM POWER PLANT ENGINEERING 



top. The reservoir is supplied with oil by the valve marked "inlet." 
When the tank is filled the oil rises in the standpipe D a corresponding 
height. The inlet valve is then closed and the oil in the standpipe 
feeds down to the level of the sight feeds or to a point where the air 
will enter the bottom of the tank. This will be the constant oil level, 
since oil flows from the tank only in proportion to the amount of air 

CHECK 



OIL TANK 



ev-PAss 



zzzzzzzzz< 






* 



nzzzz 



FLEXIBLE 

FILLING 

PIPE 



OVER FLOW 



RETURNS 



FLOOR LINE 



///////////////////// //V ^T7-TZ7 



\pl 



OIL 
RESERVOIR 



] 



I 



BASEMENT FLOOR LtNE 



Fig, 529. Simple Gravity Feed System. 

admitted. A head of 6 inches has been found to give the best results. 
(Engineer, U. S., March 16, 1903, p. 243.) 

421. Compressed-air Feed. — Fig. 531 shows diagrammatically the 
arrangement of the oiling system at the First National Bank Building, 
Chicago. The storage tank containing the supply of engine oil is 
under air pressure at all times except during the short periods when it 
is being filled with oil from the filter. The air pressure on the surface 
of the oil forces it to a manifold on the engine from which it is dis- 
tributed to the various oil cups. The oil flows from the different 



LUBRICANTS AND LUBRICATION 



763 



j£ 



5 



l F?^AIR HOLE 



<&i* 



L 



SMALL GRASS PIPE 



■A-^-a^A . 



k 



Fig. 530. Low-Pressure Gravity Feed, Constant Head. 




SETTLING TANK 



]_ 



I J 



ikj 



OIL STORAGE 



|-t> J » J - 



Fig- 531. Oiling System at the Power Plant of the First National Bank Building, Chicago 



764 



STEAM POWER PLANT ENGINEERING 



bearings to the returns tank located at the base of the engines. When 
the tank is filled air pressure is admitted and the oil forced to the settling 
tank, which has a capacity of about 400 gallons and is located near the 
ceiling. The oil is allowed to settle and the entrained water and foreign 
material are drained to waste. The oil gravitates from this tank to a 
series of Turner oil filters. When a new supply of oil is needed, valves 
A and B are closed and vent valve C opened, cutting off the supply of 
air and reducing the pressure to atmospheric. Valve D is then opened 
and oil flows from the filters to the storage tank. 

422. Cylinder Lubrication. — The test requirements for cylinder oils 
are outlined in Table 126, from which it will be seen that pure mineral 
oil fulfills practically all requirements for dry steam. In connection 
with moist steam, as in the low-pressure cylinders of compound engines, 
an addition of from 2 to 5 per cent of acidless tallow oil is recommended. 
Vegetable oils, beeswax, lard oil, degras (wool grease), and the like 
should never be used in compounding cylinder oils. The best cylinder 
oils are made from Pennsylvania stock. For data pertaining to the 
amount and grade of cylinder oil used in a large number of piston engine 
plants see Table I, p. 824, Jour. A.S.M.E., May, 1910. See also " Lubri- 
cants and Lubrication," by Dr. C. F. Mabery, Jour. A.S.M.E., Feb., 
1910. 

Cylinder oils must be forced to the parts requiring lubrication against 
the prevailing steam pressure, which is ordinarily accomplished by 
(1) cylinder cups, (2) hydrostatic lubricators, or (3) hand or power driven 
force pumps. 

423. Cylinder Cups. — A cylinder oil cup consists essentially of a 
steam-tight brass vessel fitted at the bottom with a pipe connection 

p and valve. A screwed cap offers a means of 

introducing the lubricant into the cup. After 
the cap is in place the valve is opened and the 
cup is subjected to full steam pressure. The 
pressure in the cup, being equal to that in the 
steam chest or cylinder, permits the lubricant to 
gravitate through the valve into the cylinder. 

Fig. 532 shows a section through an improved 
form of oil cup in which the oil feeds from the 
top instead of the bottom as is the case with 
the common form of cylinder cup. The vessel 
is attached to the steam chest or to the supply 
pipe below the throttle valve. Steam is admit- 
ted through opening B and, condensing, settles 
through the oil to the bottom. This raises the level of the oil until it 




Fig. 532. Leyland Auto 
matic Cylinder Cup. 



LUBRICANTS AND LUBRICATION 



765 



begins to overflow down the same passage by which the steam enters. 

This action is intensified by the fluctuation r^Ol 

in steam pressure. The rate of feeding is 

regulated by valve C and tested by un- 
screwing plug F. If oil appears through 

opening G, the cup is feeding oil; if steam 

or water is emitted the cup is empty. The 

cup is filled by means of plug E and the 

water drained at D. 
424. Hydrostatic Lubricators. — The most 

common method of cylinder lubrication is 

by means of hydrostatic lubricators of the 

sight-feed class, Fig. 533. The principle of 

operation is as follows: The lubricator is 

filled with cylinder oil by removing cap K, 

the height of oil appearing in glass L. If 

water is present the "oil floats on top as 

indicated. After the cap is screwed in place v 

the valves in the condenser pipe are opened, 

subjecting the oil in the vessel to steam- 
pipe pressure. Steam is condensed in pipe 

C, filling tube B and part of C, thus adding 

to the steam fig. 533 
pressure the 

pressure due to the weight of the water 
column. Valve F, which communicates 
with the top of the vessel by means of tube 
A, is opened wide, as is also the regulating 
valve I. The pressure at B being greater 
than that at A by an amount equivalent 
to the height of the water column, forces 
the oil through A and the " sight feed" S 
to the steam pipe. The rate of flow is 
controlled by the regulating valve I. As 
the oil flows from the vessel its space is 
occupied by condensed steam, the height 
of oil and water being visible in glass L. 
Owing to the small capacity of the lubri- 
cator it must be refilled frequently. To 
reduce the amount of labor required with 

the above apparatus, independent sight feeds, Fig. 409, are sometimes 

used in connection with a central reservoir. Such an installation is 





Common Hydrostatic 
Lubricator. 



Fig. 534. Lunkenheimer Sight- 
feed Lubricator. 



766 



STEAM POWER PLANT ENGINEERING 



shown diagrammatically in Fig. 535. A condenser pipe leading from 
the steam main enters the bottom of the reservoir and the condensed 
steam fills up the reservoir as fast as the oil is fed out. The principle 
is the same as that of the simple hydrostatic lubricator. Oil is fre- 
quently injected by mechanical means under a steady pressure gen- 
erated and governed independently of the steam. Two systems are 
in common use, direct mechanical pump pressure and air pressure. 

425. Forced-feed Cylinder Lubrication. — Fig. 536 illustrates the 
" Rochester" simple feed automatic lubricating pump, which takes the 
oil by gravity from the reservoir through a sight-feed glass and forces 
it through a small pipe to the steam supply pipe. The pump entirely 
obviates the trouble due to intermittent feeding and, being directly 



Steam Main 






,K 



? 



iHfrfl 



T=T 




CQH 



1 r 



CZZJ 




o 



To Other Engines 



Cylinder Oil 
Reservoirs 



Fig. 535. Central Hydrostatic Lubricator. 

driven from the engine, runs at constant speed. The feed is uniform 
and independent of the pressure pumped against. The rate is deter- 
mined by the length of stroke of the pump piston, which is easily 
adjusted. 

With large engines multi-feed pumps are sometimes used, which force 
oil to the various valves as well as to the steam pipe. Fig. 537 shows 
an arrangement of storage tank in connection with pump reservoir to 
avoid the trouble of hand filling. 

426. Central Systems. — Fig. 538 shows the piping for a large central 
system of cylinder and engine lubrication. There are two storage tanks 
on the engine-room floor, one for cylinder oil and the other for engine 
oil, the distributing arrangements being the same in each case. The 
oil is pumped from each tank into a main pipe extending the length of 
the engine room and provided with branches at each point requiring 
lubrication. The oil pumps are actuated by steam and are of the duplex 
direct-acting type, provided with automatic governors which regulate 



LUBRICANTS AND LUBRICATION 



767 



the speed to suit the demand for oil. The cylinder oil is forced through 
a special sight-feed lubricator, Fig. 539, under a pressure of about 
25 pounds in excess of the steam pressure. Referring to Fig. 539, 
diaphragm valve D, in the bottom of the lubricator, is kept closed by 




Fig. 536. Rochester Forced-feed Lubricator. 

the steam pressure admitted through pipes B. Thus the inlet pressure 
must be greater than that of the steam before the valve will open and 
admit oil to the engine. The oil, after entering, passes upward through 
the sight-feed glass and downward through the hollow arm A to the 



H.P.Steam Pipe 

L.P. Steam Pipe 
To Rod \ 



To Rod 




Fig. 537. Forced-feed Cylinder Lubrication. 

steam pipe. The engine oil is forced by the pump to the various points 
under a pressure of about 20 pounds. The waste oil is caught in suit- 
able receptacles and, after being filtered, is returned to the storage tank 
by a steam pump. This pump is connected so that it can supply the 
storage tank either from the filter or with fresh oil from a large oil tank 



768 



STEAM POWER PLANT ENGINEERING 





f— *~lf£-\ — ■ 

_ s-bHr 1 — J 

r- 






A-^--4 



L-lOiO LJ 







LUBRICANTS AND LUBRICATION 



769 



in the basement. By this arrangement all handling of oil in the engine 
room is done away with. 

427. Oil Filters. — After oil has been applied to machinery its lubri- 
cating properties become impaired on account of (1) contamination 



j^i 




TO STEAM 



FEED REGULATOR 



Fig. 539. Siegrist Sight-feed Lubricator. 



with anti-lubricating material, such 

as dust, metallic particles from wear, 

gum, acid, and resin; and (2) ex- 
posure to heat and the atmosphere 

which drives off part of the more 

volatile constituents and decreases 

the fluidity of the oil. 

In many small plants no attempt 

is made to reclaim oil that has once 

been used, since the quantity is so 

small that the cost and trouble 

involved would more than offset 

the gain. Where large quantities 

of oil are used, considerable saving 

may be effected by using it over and over again. To render the oil fit 

for reuse it must be thoroughly purified. The anti-lubricating matter 

is removed by precipitation and filtration. 

Fig. 540 shows a section through a " White Star" oil filter and purifier. 

The apparatus consists of a cylindrical sheet-iron vessel divided into 

two compartments by a ver- 
tical partition. These two 
compartments are con- 
nected near the top by valve 
B. The smaller chamber is 
provided with a funnel A 
and a steam coil for heating 
the contents. The large 
chamber contains a cylin- 
drical wire screen covered 
with several folds of filtering 
cloth. Impure oil is poured 
into funnel A , the upper part 
of which is provided with a 
removable sieve or strainer, 

and is discharged below the surface of the water through holes in the foot 

of the tube. The thin streams of oil rise vertically to the surface of the 

water and the heavy particles of grit and dirt gravitate to the bottom. 

The steam coil heats the oil and water and facilitates precipitation of 




WATER LEVEL 



Fig. 540. 



\ DRAIN 

White Star Oil Filter. 



770 



STEAM POWER PLANT ENGINEERING 



the solid matter by thinning out the streams of oil. When the oil in the 
smaller chamber reaches the level of valve B it flows into the filter bag, 
which removes the remaining impurities and permits the purified prod- 
ucts to flow into the large compartment from which it may be drawn 
at will. All parts are accessible and readily removed for cleaning pur- 
poses. The accumulated sediment in the bottom of the small chamber 
is discharged to waste at intervals by means of a suitable drain. When 
the filter cloth is to be removed, valve B is closed and the wire cylinder 
is disconnected and lifted out. Any oil remaining in the filter is re- 
turned to funnel A . The filter cloth is held against the screen by cords 
and hence is readily removed. 



-Perforated Plate 

Water 

Steam Coils 




SECTION 



SECTION 2 



SECTION 3 



SECTION 4 



Fig. 541. Turner Oil Filter. 



Fig. 541 shows a section through a Turner oil filter, illustrating the 
type of filter usually installed in large stations where continuous fil- 
tration is desired. This apparatus consists of a rectangular tank 
divided into four compartments. The returns from the lubricating 
system flow into section 1 through a screened funnel and discharge 
into the water space at the bottom of the compartment. The oil rises 
through the water, passes, under pressure of the head in the funnel, 
through a layer of filtering material resting on a perforated plate, and 
collects in an inverted cone. Through perforations around the top of 
the cone it passes into a dirt chamber, where most of the heavy impuri- 
ties are deposited, and then, still rising, passes through another per- 
forated plate and more filtering material. The partially cleaned oil, 
which issues, overflows into the second compartment and thence into 
the third, the same cycle of operations being repeated in these two. 
The overflow from the third compartment descends through a final 



LUBRICANTS AND LUBRICATION 771 

filter in the fourth compartment and collects at the bottom, from which 
it is withdrawn by the oil pump. 

Cylinder Lubrication: Power, Feb. 15, 1910, Jan., 1905, p. 36; Engr., Lond., 
1903, Vol. 96, pp. 55, 108, 132, 155; Engr. U. S., Oct. 15, 1906, p. 682; Jour. A.S.M.E., 
Feb. and May, 1910. 

Miscellaneous. — Measurement of Durability of Lubricants: Trans. A.S.M.E., 
11-1013. Valuation of Lubricant by Consumer: Trans. A.S.M.E., 6-437. Suit- 
ability of Lubricants: Power, Nov., 1906, p. 673. Oil Required for Lubricators: 
Elec. World, May 5, 1906, p. 934. Gumming Tests: Jour. Am. Chem. Soc, April, 

1902, p. 467. Valuation of Lubricants: Jour. Soc. Chem. Ind., April 15, 1905, 
p. 315. 

Lubrication, General: Power, Sept. 12, 1911, p. 396; Prac. Engr., Feb. 15, 1912, 
p. 194. 

Oil Purification: Elec. World, Dec. 1, 1906, p. 1053. 

Economy in Lubrication of Machinery: Trans. A.S.M.E., 4-315. Theory of 
Finance of Lubrication: Trans. A.S.M.E., 6-437. 

Experiments, Formulas, and Constants for Lubrication of Bearings : Am. Macfr., 

1903, pp. 1281, 1316, 1350. 

Lubricators and Lubricants: Power, Sept. 21, 1909, p. 486; Feb. 22, 1910, p. 347. 
Selection of an Oil for Lubrication: Power, July 27, 1909, p. 137. 



CHAPTER XVII. 



TESTING AND MEASURING APPARATUS. 



428. General. — The importance of maintaining a system of records 
is discussed in paragraph 459. The various items which may be re- 
corded and the instruments and appliances used in this connection are 
outlined in the accompanying chart. In large stations a full comple- 
ment of indicating, recording, and integrating instruments may prove 
to be a good investment if intelligently and closely studied by the 
operating engineer with a view to locating and eliminating unnecessary 
losses. The instruments should be inspected and calibrated at inter- 
vals, since many of them are delicately constructed and are apt to 
become inaccurate after a few months' service. Steam gauges, ther- 
mometers, and pyrometers, and particularly water meters are subject 
to appreciable error after considerable use. Voltmeters, ammeters, and 
other switchboard instruments are easily deranged, especially when 
subjected to continuous vibration or to high temperature. 

429. Weighing the Fuel. — In most small plants the 
delivery tickets of the coal dealer are depended upon 
for the weight of coal used, no attempt being made 
to determine its evaporative value, and the economy 
of the plant is judged by the size of the coal bill. In 
such cases a considerable saving may be /effected by 
keeping a daily record covering at least the coal and 
water consumption. The coal can be conveniently 
weighed on ordinary platform scales. In a number 
of large stations the weight of coal is determined by 
suspended weighing hoppers, which may be station- 
ary, as in Fig. 148, or mounted on^a traveling truck, 
as in Fig. 149. The scales of such devices are made 
indicating, autographic, integrating, or a combination 
of the three, the latter costing but little more than 
the simple indicating or recording devices. 

A simple and inexpensive coal meter recently 
brought out is illustrated in Fig. 542. It consists 
Fig. 542. Coal Meter, essentially of a helical vane placed in a cylindrical 
conduit. The movement of the coal causes the vane to rotate and 
the number of revolutions is a measure of the weight of fuel pass- 

772 




'-• ■ • ■ '-■:^-i---4 . .. 



— _ 



TESTING AND MEASURING APPARATUS 



773 



TESTING AND MEASURING APPARATUS. 



Tempera- 
tures 



Power . 



(Fuel 



Weights s Water 



Pressures . . . 



Flue gas 
analysis 



Moisture. . 



Fuel analysis 



Steam . 



STEAM PLANT. 

(Platform scales, indicating and autographic. 
Suspension hoppers, indicating and auto- 
graphic. 
Coal meters, integrating. 
Platform scales and tanks. 

Piston. . . . 1 

Rotary. . . . [ Indicating. 

Water meters . -( Disk J 

Venturi, indicating and au- 
tographic. 
Weirs and volume displacement meters. 
Weighing condensed steam. 
Direct. 



Steam meters 



' Hih C Bourdon gauge 

" I Manometers, mercurial, indicating 

J Manometers — 
autographic. 
Manometers — 
graphic. 



Indirect. 

indicating and autographic, 
ircurial, indicating, 
mercurial, indicating, and 

water, indicating, and auto- 



800 to 2500 deg. F. 



Over 2500 deg. F... 



V Diaphragms, indicating and autographic. 
Mercurial thermometers, indicating. 
Up to 800 deg. F. . . ^ Expansion thermometers, indicating and 
autographic. 
Expansion thermometers, indicating and 

autographic. 
Resistance thermometers, indicating and 

autographic. 
Thermo-electric thermometers, indicating 
and autographic. 
( Optical pyrometer, indicating and auto- 
< graphic. 

( Platinum or clay ball pyrometer, 
dicated I Indicators, hand manipulated. 

' " I Indicators, continuous autographic, 
f Rope brake. 
nmmin»^ J Prony brake, 

developed -j Absol p t i on dynamometers. 

I. Electric generator. 
r Orsat apparatus. 
J Arndt's econometer, indicating, 
j Westover recorder, autographic. 
(.Uehling gas composimeter, autographic. 
( In air Hygrometer, indicating and autographic. 

j In steam Calorimeters . . j Ikrotth^f .' 

f Mahler bomb. 

Coal calorimeters... ?—. 

I Parr. 
Gas calorimeter Junker. 



ELECTRICAL PLANT. 

Voltage Voltmeters, A. C. and D. C, indicating and autographic. 

Current Ammeters, A. C. and D. C, indicating and autographic. 

Output Wattmeters, A. C. and D. C., integrating and autographic. 

Power factor. . .Power factor meters, A. C. only, indicating and autographic. 

Frequency Frequency meter, A. C. only, indicating. 

Synchronism. . . Synchronizers, A. C. only, indicating. 



774 STEAM POWER PLANT ENGINEERING 

ing. For hard coal of uniform size the meter gives consistent results 
agreeing within two per cent of scale weight, but with bituminous 
coal the results are somewhat erratic and particularly so with lumps 
of varying size. (For a detailed description of the device, see Prac. 
Engr., U. S., Apr. 15, 1912, p. 438.) 

430. Measurement of Feed Water. — The quantity of water fed to 
the boiler may be determined by 

1. Actual weighing. 

2. Measurement of volume displacement. 

3. Measurements by weirs and orifices. 

4. Measurement by determining the velocity of flow in the feed pipe. 
Some of these methods necessitate measurement on the suction side 

of the pump; others are applicable to either suction or pressure. The 
former, as a class, are the more accurate but involve bulky apparatus. 
The choice for any given case depends upon the quantity of liquid to 
be measured, the degree of accuracy required, space requirements, and 
first cost. 

431. Actual Weighing of Feed Water. — The most accurate means of 
measurement is by the use of two or more tanks resting upon scales, 
arranged to be filled and emptied alternately. This method is limited 
to comparatively small quantities because of the great bulk of apparatus 
involved and is seldom used for continuous service. It is commonly 
employed in conducting special tests of short duration and for calibra- 
tion purposes. For regular boiler service it involves considerably more 
time than is ordinarily at the disposal of the fireman and engineer. 
For temperatures above 150 degrees F., the weighing tanks should be 
covered, since evaporation may cause an appreciable error. See also 
" Rules for Conducting Boiler Trials," A.S.M.E., Code of 1912, re- 
printed in Appendix B. 

432. Worthington Weight Determinator. — Fig. 543 shows the gen- 
eral details of the Worthington weight determinator, illustrating a 
commercial means of continuously measuring and recording the weight 
of water fed to the boiler. The apparatus consists primarily of two* 
tanks of equal size, A and B, each mounted on knife edges K and 
equipped at one end with a siphon S and at the other end with counter 
weight W. The liquid to be measured flows through inlet pipe P and 
along deflector D into either tank. Each tank remains in a horizontal 
position until the weight of liquid overcomes the counter weight when 
it tilts into the position shown by the dotted lines. Discharge now 
takes place through siphon S until the liquid reaches a certain level at 
which point the tank tilts back to its original position and the siphon 
continues its action until the vessel is emptied. The tanks operate 



TESTING AND MEASURING APPARATUS 



775 



alternately, one filling while the other is discharging. Since each tilt 
represents a definite weight of liquid irrespective of variations in volume 
due to specific gravity or changes in temperature, the number of tilts 
as recorded by counter C is a correct measure of the weight discharged. 
This apparatus operates at atmospheric pressure and is arranged to 
discharge into a storage tank from which the feed pump takes its supply. 

X 




hW 



SECTION X-Y 

Fig. 543. Worthington Water Weigher. 

433. Kennieott Water Weigher. — This apparatus is used in many 
boiler houses and seems to give universal satisfaction. It consists of a 
cylindrical shell S, Fig. 544, the lower part of which is divided into two 
measuring compartments A and B, each fitted with a siphon for dis- 
charge and a float F for actuating the tripping mechanism. Tripping 
box E is divided into two sections which alternately fill with water and 
serves the double purpose of furnishing a sufficient quantity of water 
to start the siphons and to shift the supply from one compartment to 
the other. This tripping box is balanced on knife edges and is mounted 
directly above the measuring compartments. Water enters the inlet 
and passes to the tripping box where a small portion is intercepted, 
the remainder passing directly to the measuring compartment below. 
When this compartment is nearly filled the float tilts the tripping box, 
discharges its contents into the compartment, and starts the siphon. A 
counter registers each double charge. This apparatus discharges at 
atmospheric pressure, though with slight modification it may be in- 
stalled on the pressure side of the pump. Kennieott water weighers 
are constructed in various sizes ranging from a capacity of 750 to one 
million pounds per hour and are guaranteed by the manufacturers to 
record the correct weight of water within one-half of one per cent of 
scale weight at any given temperature. Calibration for different tem- 
peratures is necessary since the apparatus is actuated by volume dis- 
placement. For example, the weight of one cubic foot of water at 



776 



STEAM POWER PLANT ENGINEERING 



60 degrees F. is 62.37 pounds and at 210 degrees F. it is 59.88, a differ- 
ence of 2.49 pounds. Hence, if the device is calibrated to read correctly 
at 60 degrees it would be in error 4 per cent if used to measure water 
at 210 degrees F. 

434. Willcox Water Weigher. — Another successful volume displace- 
ment meter is illustrated in Fig. 545. The device consists of a cylin- 
drical tank divided into an upper and lower compartment by a hori- 
zontal partition. The water enters the upper compartment, passes to 
the lower, in which its volume is measured, and then out through the 
U-shaped discharge pipe. The operation, beginning with both com- 
partments empty, is as follows: Water enters the upper compartment 




Discharge Pipe 



Fig. 544. Kennicott Water Weigher. 



Fig. 545. Willcox Water Weigher. 



through the inlet pipe and rises to the top of the standpipe. (The latter 
is open at the top and bottom and is rigidly connected to the bell float, 
but when in its lowest position it is held against its seat by weight of 
the bell float.) Further admission of water causes it to overflow into 
and through the standpipe into the lower compartment. The water, ris- 
ing in the lower compartment, seals the lower edge of the bell float and 
entraps a volume of air under the bell. Further rise compresses the air 
under the float, in leg C of the discharge pipe and in leg A of the trip 
pipe AB. This compression causes the float to rise to its highest posi- 
tion and raises the standpipe from its seat, permitting the water in the 
upper chamber to pour into the lower vessel. Compression of air con- 
tinues until the pressure becomes great enough to break the seal in the 
trip pipe. This action immediately reduces the pressure below the 
float, permits the latter to descend, sealing the upper chamber against 



TESTING AND MEASURING APPARATUS 



777 



further discharge, and allows the water in the lower compartment to 
siphon out through the discharge pipe. The number of discharges is 
recorded mechanically. 

435. Weir Measuring Devices. — Feed water heaters or specially 
designed tanks fitted with V-shaped or trapezoidal weir notches offer 
a simple means of measuring the instantaneous rate of flow. The 
heater chamber is divided into vertical compartments arranged so that 
one may discharge through a calibrated weir notch into the other. 
The height of water above the bottom of the notch is a direct measure 
of the volume flowing. The height may be noted in an ordinary gauge 
glass or it may be transferred through a suitable float mechanism to an 
outside indicator. In determining the total flow for any given period 




Fig. 546. A Typical Piston Water Meter. (Worthington.) 

it is necessary to take readings at frequent intervals. The chief draw- 
back to this method lies in the difficulty of preventing sudden fluctua- 
tions in water level due to surging in the heater. For the theory of 
weir notches, orifices, and nozzles consult " Experimental Engineering," 
Carpenter and Diederichs, 1911, Chapter XII. See also, Jour. A.S.M.E., 
Oct., 1912, p. 1479. 

436. Pressure Water Meters. — There are a number of reliable water 
meters on the market for hot or cold water which may be placed on the 
pressure side of the feed pump. Among them may be mentioned the 
Hersey, Crown, Nash, and Worthington. They are all based on volume 
displacement and consequently require correction for different tem- 
peratures if graduated to read in pounds. They are compact, com- 
paratively inexpensive, and require considerably less space than the 
tank weighers of the Kennicott and Willcox type but are open to the 
objection that no particular provision is made against leakage and after 
considerable use they are subject to serious error. In many plants where 
meters of this type are installed the meter is by-passed and operated 
only for short periods. For continuous service meters of the tank- 



778 



STEAM POWER PLANT ENGINEERING 



weighing or Venturi type are recommended. Fig. 546 illustrates the 
piston type of pressure meter, in which reciprocating pistons are dis- 
placed by a definite volume of water; Fig. 420, the rotary type, de- 
pending upon the displacement of rotary impellers; Fig. 547, the disk 
type, in which impellers are given a combined rotating and tilting 




Fig. 547. 



A Typical Disk Water Meter. (Nash.) 



motion. The capacities of pressure meters, irrespective of type or 
maker, range approximately as follows: 

Size of meter (pipe size) f , 

Maximum capacity, cubic feet per minute : 

Rotary or disk meters 1, 

Piston meters 



1 

2, 


3 

4j 


1, 11, 2, 


3, 


4, 


6 


2, 


4, 


8, 12, 20, 


36, 


72, 


120 


i 

2) 


3, 


5, 6, 8, 


23, 


60, 


120. 



437. Venturi Meter. — The Venturi tube with indicating, auto- 
graphic, and integrating mechanism, as constructed by the Builder's 
Iron Foundry of Providence, R. I., is one of the most satisfactory 
methods of measuring feed water under pressure. The total absence 
of working parts in the meter proper insures continuity of operation 
and freedom from wear, and the fact that the recording mechanism 
may be placed at a considerable distance from the meter is a great 
advantage. The Venturi tube, Fig. 548, is essentially the same in 
principle as an orifice placed in the pipe. The pressure difference H 
between A in the " upstream" portion of the tube and B at the " throat" 
is a measure of the velocity through the throat. The loss of head due 
to friction is negligible and the velocity may be calculated, within an 






TESTING AND MEASURING APPARATUS 



779 



error of 3 per cent, from the following modification of Bernouilli's 
theorem : 



VF U 2 - F t 2 
in which 

V t — velocity at the throat, feet per second. 
F u = area of the upstream section, square feet. 
F t = area of the throat, square feet. 
H = pressure difference, feet of water. 



(284) 



vuh. 



Inlet 



Pipes to Manometer 



— H^ 



HP^iRIF 



Outlet 



I ^jn 



Fig. 548. Venturi Tube with Indicating Manometer. 



Manometer 



For accurate work the tube requires calibration. Once calibrated the 
error in weight readings for a given temperature should not exceed one 
per cent for capacities within the working range of the manometer. 
For very low throat velocities the error may be considerable because 
of the slight pressure difference between the points A and B. In situa- 
tions where there are periods of very low and very high rates of flow, 
as in connection with combined heating and lighting plants, it is cus- 
tomary to install a small tube for the light loads and a large tube for 
the heavy loads, the same indicating mechanism being used in each 
case. The equipment illustrated in Fig. 548 is purely indicating and 
readings must be taken at frequent intervals in order to obtain the total 
flow for a given period. Where the size of the plant warrants the out- 
lay the combined indicating, integrating, and recording instrument is 
often installed. With this device the instantaneous rate of flow is 
indicated by a pointer and dial, the variation in rate of flow for any 
given period is recorded on a clock-driven chart, and the total weight 
flowing is registered on a counter. (For a detailed description of this 
mechanism see Power, Jan. 23, 1912, p. 102.) Tests made at Armour 
Institute of Technology on a carefully calibrated tube and recorder 
with feed water at 210 degrees F. and constant rate of flow gave chart 



780 



STEAM POWER PLANT ENGINEERING 



and counter readings agreeing substantially with scale weights; for 
irregular and fluctuating flow, as when feeding the boilers, the average 
error was about one per cent. 

438. Orifice Measurements. — The appropriation of the great majority 
of small steam power plants does not permit of the installation of 
tank meters, Venturi meters, or other forms of reliable commercial 
appliances for measuring the weight of water fed to the boilers. For 
use in such cases an inexpensive and fairly accurate indicating meter 
may be constructed of ordinary pipe fittings, as illustrated in Fig. 549. 
A thin metal diaphragm with circular orifice is inserted on the pres- 
sure side of the feed pump and the pressure drop across the orifice is 
measured by inclined mercury manometer. The height of mercury 



Bypass Union. 




vllron Gage Cock 

Fig. 549. Simple Indicating Water Meter, Orifice Type. 



h is an indication of the rate of flow. By calibrating the manometer 
against tank measurements the readings of the mercury column may 
be graduated to read directly in pounds per hour. If means are not 
available for calibration purposes the weight of discharge may be 
approximated from the formula 

W = 1120 a Vhd, (285) 

in which 

W = weight flowing, pounds per hour. 
a = area of the orifice, square inches. 
h = vertical height of mercury column, inches. 
d = density of the water, pounds per cubic foot. 

For a fairly continuous flow and pressure drop corresponding to three 
inches of mercury or more this simple device gives results agreeing 
within four per cent of tank weights, but for widely fluctuating flow 
and small pressure drops the error may be considerably more. 



TESTING AND MEASURING APPARATUS 781 

For application of the Pitot tube for water measurements consult 
accompanying bibliography. 

The Pitot Tube for Water Measurements: Trans. A.S.M.E., Vol. 30, 1908, p. 351; 
Vol. 25, 1904, p. 184; Vol. 22, 1901, p. 284. 

The Pitometer: Proc. Am. Water Wks. Asso., 1907, p. 136; Jour. Frank. Inst., 
Dec, 1907, p. 425. 

439. Measurement of Steam. — The quantity of steam passing 
through any device may be determined by (1) condensing and weighing 
the steam after it has passed through the apparatus and by (2) measur- 
ing the flow by means of steam meters before it enters. The first 
necessitates the use of surface condensers, and consequently has a 
limited field of application, whereas the latter may be used in both 
condensing and non-condensing service. 

440. Weighing Condensed Steam. — The weight of condensed steam 
may be obtained by any of the devices used in connection with feed 
water measurements but such measurements are seldom made except 
for test purposes because of the expense or labor involved. The Wheeler 
Condenser and Engineering Company's " indicating hot well" offers a 
practical and simple solution of continuously measuring the condensed 
steam. The hot well is attached to the bottom of the condenser cham- 
ber in the usual way and differs from the ordinary hot well only in the 
addition of a vertical partition. This partition divides the hot well 
chamber into two compartments. Condensation from the condenser 
drains into one of these compartments and flows to the other through 
a calibrated orifice. The height of water above the orifice as shown in 
the gauge glass is an indication of the weight of condensation flowing. 
The manufacturers guarantee an accuracy within 2 per cent of scale 
weight for readings over the whole range. The readings are purely 
indicating and must be taken at frequent intervals in order to give the 
total flow. 

441. Steam Meters. — The weight of fluid flowing through an open- 
ing may be calculated by the equation 

W = AyV, (286) 

in which 

W = weight in pounds per second. 

A = cross-sectional area in square feet. 

y = density of the fluid, pounds per cubic foot. 

V = velocity of flow, feet per second. 

All steam meters for indicating or recording the weight of steam 
flowing through a pipe are based upon the law expressed in equation 
(286). Thus, for steam of constant density the opening through which 
it flows may be made constant and the variation in velocity will be an 



782 



STEAM POWER PLANT ENGINEERING 



indication of the rate of discharge; or the velocity may be held constant 
and a variation in the amount of opening will be an indication of the 
weight discharged. Unfortunately, the density of steam is seldom 
constant under commercial conditions and herein lies the inherent 
defect of all steam meters which depend for their operation upon a 
variation in the area of efflux or a variation in velocity. The density 
of steam is a function of its pressure and quality and any variation in 
either will affect the weight of discharge as determined from equation 
(286). Pressure variations may be automatically compensated for, but 
corrections for quality must be made in each specific case. 



CLASSIFICATION OF STEAM METERS. 

Lindenheim (1896)* 
Gebhardt (1908) f 



Indirect Velocity 



Direct 



Throttling 



Pitot tube 



< Current 



f Floating 
valve 



Stationary 
disk 



Venturi 
tube 



Impeller 



Water ( Burnham (1905) f 

manometer ( Clyde (1910) f 



Mercury 
manometer 



Impeller 



Mechanical 
control 



5 Mercury 

I manometer 

Bourdon 
manometer 



General 

Electric (1910)*f 



Holly (1877)* 



[ St. Johns (1893)ft 
Gehre (1896)tt 
Baeyer (1902)tt 
Bendemen (1902) ft 
Sargent (1908) f 
Lindmarkft 

( Gehre-Hallwachs 
J (1907-1910)*tt 
1 Sarco (1910)*tt 
tStorrer (1910) ft 

| Eckardts (1903) tt 



Mercury f Parenty (1886) f J 

manometer \ Builders' Iron 

[ Foundry (1910)ft 

The different means adopted for transmitting this area and velocity 
variation to the indicating or recording devices overlap to such an 
extent as to render a classification of steam meters very unsatisfactory. 
The accompanying chart is offered as a guide in grouping the most 
commonly known devices. From this chart it will be seen that all 
meters may be grouped into general classes, direct and indirect. The 
direct meter is an integral part of the piping and the entire mass of 
* Integrating. f Indicating. J Autographic. 



TESTING AND MEASURING APPARATUS 



783 



fluid to be measured passes through the apparatus. It is not portable 
and cannot be readily applied to pipes of different sizes. In the indirect 
meter only a small part of the fluid to be measured is directed through 
the apparatus and the pipe line need not be disconnected for its installa- 
tion. One instrument suitably calibrated may answer for any size of 
pipe. 

The average high-grade steam meter is a reliable and accurate means 
of measuring the flow of steam in straight lengths of pipes, provided 
the flow is continuous or that the change in the rate of flow is gradual and 
the pressure and quality are practically constant. For interrupted or 
intermittent flow and for sudden variations in pressure or quality, the 
results are not reliable and may be considerably in error. The accuracy 
of all meters, provided they have been correctly calibrated and adjusted, 
depends largely upon the degree of refinement in reading the indicators 
and in integrating the charts. The commercial failure of many steam 
meters is due to the fact that they are not cared for or operated in 
strict accordance with the principles of design. 

Only a few of the best-known meters will be described here. For a 
detailed discussion of the various types of steam meters see the author's 
paper " Various Types of Steam Meters," Power, Feb. 6 and 13, 1912. 




Figs. 550, 551, 552. Principles of the " Gebhardt" Indicating Steam Meters. 

442. " Gebhardt " Steam Meters. — Figs. 550 to 554 illustrate various 
forms of indicating steam meters designed and tested at the Armour 
Institute of Technology, which are based on the principles of the 
Pitot tube. Referring to Fig. 550, A and C are two ordinary gauge 
cocks and G is a common gauge glass, C being connected with the static 
nozzle S and A with the dynamic tube D. The height of water H is 
proportional to the square of the velocity of steam flowing through 
pipe P and automatically adjusts itself to the variations in velocity; 
thus, for decreasing velocities, the water in glass G discharges through 



784 STEAM POWER PLANT ENGINEERING 

D until the water column H balances the velocity pressure in pipe P t 
and for increasing velocities, condensation from the upper part of the 
instrument accumulates and the water column H rises until a balance 
is effected for the higher velocities. 

The relation between the height of the water column and the velocity 
of the steam in the main pipe at the entrance to the dynamic tube may 
be determined from the well-known equation 

V = cV2ih, (287) 

in which 

V = maximum velocity of flow, feet per second. 

c = coefficient determined by experiment. 

h = height of a column of steam equal in weight to the water 

column H. 

This equation may be expressed 



k\Jh 



in which * d a 

K = coefficient determined by experiment. 
H = height of water column in inches. 

d w = density of water in gauge glass, pounds per cubic foot. 
d s = density of steam in the main pipe. 

Because of the labor of determining the relationship between the 
mean and the maximum velocity for various conditions of flow and dif- 
ferent pipe diameters it is more satisfactory to calibrate the gauge, by 
actual experiment, to read directly in pounds per hour. 

This simple device in connection with a calibrated scale gives readings 
within 5 per cent of condenser measurements for continuous flow and 
constant pressure and quality of steam (for velocity pressures corre- 
sponding to lj inch of water or more). For a considerable variation in 
pressure and quality or for marked changes in rate of flow the instru- 
ment is not reliable. Its sensitiveness is greater at high velocities, since 
the height of water column in the gauge glass increases with the square 
of the velocity of the steam in the main pipe. For interrupted flow, as 
when connected to a high-speed engine, the water column may be made 
to closely approximate the mean velocity by suitably throttling the 
gauge cocks. 

Figs. 551 and 553 show application of the same principle with only 
one connection to the main pipe. The commercial meter (Fig. 553) 
gives readings within 2 per cent of condenser weights for velocity 
pressures corresponding to 1 inch of water or more. Fig. 552 shows 
another form which may be placed below or above the point in the main 
pipe at which the Pitot tubes are placed. The operation is as follows: 



TESTING AND MEASURING APPARATUS 



785 




Fig. 553. 



Commercial Form of "Gebhardt" 
Steam Meter. 



Velocity pressure is transmitted through tube D and opening 0, into the 
body of the chamber M . This pressure, acting on the surface of the 

——- ' n 



ber, forces the water into the 
glass W until a balance is ef- 
fected. Condensation is dis- 
charged continuously through 
pipe P and the water seal U 
of the main pipe. Tests of 
this meter have given results 
agreeing within 2 per cent of 
condenser measurements for 
continuous flow for all veloc- 
ities ranging from the equiva- 
lent of a 1-inch to a 10-inch 
water column. No provision 
is made for automatic correc- 
tion of pressure and quality 
variation in any of these de- 
vices. (For the theory and 
results of tests of the Pitot type of steam meter see author's paper "The 
Pitot Tube as a Steam Meter," Trans. A.S.M.E., Vol. 31, p. 603.) 

Fig. 554 shows a modified form for very low velocities or for low- 
density steam. The rise and fall of the water column is transmitted 
through a small float and multiplying gears to an indicating dial and 
pointer. Referring to the illustration, S is the static tube and D the 
dynamic tube. The movement of float B is transmitted through levers 
L and N to sector G and pinion P. Pinion P is secured to bar magnet 
M, so that the rise and fall of float B cause the magnet to rotate. Oppo- 
site magnet M, but outside the casing, is another bar magnet M r (see 
figure in lower corner of illustration). The movement of bar M within 
the casing is transmitted to bar M f outside the casing, magnetically, 
thus obviating the use of stuffing boxes. 

443. General Electric Steam Meters. — The General Electric Com- 
pany has recently placed on the market a number of steam meters of 
the Pitot-tube type in which a mercury column is used to measure the 
velocity pressure. The principle involved is an old one, but this com- 
pany is the first to exploit it successfully from a commercial standpoint. 
Three styles are manufactured, (1) in which the velocity pressure is 
measured directly by means of a U-tube manometer; (2) in which the 
variation in the height of the mercury is transmitted to an indicating 
dial through the agency of floats and pulleys; and (3) in which the 



786 



STEAM POWER PLANT ENGINEERING 



variation in the weight of the mercury column actuates a recording 
mechanism by means of a series of compound levers. 




Fig. 554. "Gebhardt" Steam Meter for Low Velocities. 

The principle of the simple indicating device is illustrated in Fig. 555, 
in which S is the static nozzle at right angles to and D the dynamic 

nozzle facing the current; U is an ordinary 
U-tube manometer partially filled with 
mercury. When there is no flow the sur- 
face of the mercury in the columns N and 
W will be on the same level and the upper 
portions will be filled with condensed vapor. 
When there is a flow, the mercury will be 
depressed as indicated and the difference H 
in the heights of the mercury columns will 
be a measure of the velocity of flow at the 
point in the pipe where the dynamic tube 
is placed. 

This velocity may be expressed by sub- 
stituting the proper values in equation (287) , 

Cirn 





fS— 






" k ? 




\ "|Sp~*) 










1 1 


: 


1 


1 






It 




: 


: 






- 


: :: 


III 


NJ 


| 


111 

W 


III 




jfj 1 


u 





Fig. 555. Pitot Tube with 
Mercury Manometer. 



thus 



= K\J) 



(289) 



TESTING AND MEASURING APPARATUS 



787 



To ".Leading So* 



in which 

d m = density of the mercury in pounds per cubic foot. (Other nota- 
tions as in equation (288)). 

A comparison of equations (288) and (289) will show that the mercury 
manometer is less sensitive than the water manometer by an amount 
equivalent to d m -s- d w or approximately 13.6. The variable height of 
the water column above the mercury is usually included in the value 
of the coefficient K. 

The General-Electric indicating-flow meter, 
shown in Fig. 556, differs from the simple 
device just described in that a simple nozzle 
plug, shown in detail in Fig. 557, is used in 
place of the ordinary static and dynamic 
nozzles. Referring to Fig. 557, TT are the 
static openings or " trailing set" and LL the 
dynamic openings or "leading set." The 
plug is screwed into the pipe with the " leading 
set" directly facing the current and the con- 
nections to the manometer are made through 
the openings T' and L' '. The weight of steam 
flowing may be obtained directly from the 
height of the mercury column H, Fig. 556, by 
means of suitable charts based upon experi- 
ments. Adjustment for variations in pressure, Fig. 556. General Principles 

quality, and pipe diameter are made by of the G.-E. indicating-flow 

setting the chart cylinder C in accordance Meter - 

with the graduated scales at the bottom of the instrument. 

For general purposes a single revolving chart is furnished, the readings 
of which, multiplied by the area of the pipe, give the weight of steam 



Plugged 



Fig. 557. Nozzle Plug; G.-E. Steam Meter. 

flowing. For low velocities the difference in the heights of the mercury 
columns, if vertical, is so small as to lead to serious error; hence, pro- 




for Quality 



mjjKAdjustment for 
-* Pipe Diameter 



I ' ' i ^ Adjustment? 
Adjustment for £or PreS81 «e 
Height of Chart 








= 
















fe 




a , 








= 


















L-i 






1 
























: 




















788 STEAM POWER PLANT ENGINEERING 

vision is made for this by inclining the manometer as indicated in 
Fig. 558. With this the actual head of mercury due to the velocity is 
H, but the difference in the lengths of the columns is D. The indication 

on the chart corresponding to the 
height of the mercury in the glass T" 
multiplied by a constant depending 
upon the inclination of the chart is 
the rate of flow in pounds per hour per 
square inch of the pipe cross-section. 
The accuracy of this meter de- 
pends entirely upon the refinement of 
Fig. 558. Inclined Mercury adjustment and the extent of error 
Manometer. j n reac ling the height of the mercury 

column. If correctly set, an error in reading of T \ inch for a velocity 
corresponding to a vertical column J inch high would be 12 J per cent, 
whereas the same error referred to a vertical column 6J inches high 
would be but 1 per cent. Tests of this instrument, conducted at 
the Armour Institute of Technology, gave readings for continuous flow 
agreeing within 1 to 8 per cent of condenser measurements, depending 
upon the rate of flow. For interrupted flow the departure from con- 
denser readings was more marked. This meter is portable, simple in 
construction, and readily applied to any pipe by inserting the nozzle 
plug at the required point, although considerable care is necessary in 
setting it up and in making the necessary adjustments for pipe diameter, 
steam pressure, and quality. 

The G.-E. Boiler-flow meter differs from the one just described in that 
the variation in height of the mercury column is transmitted through 
a float and pulleys to two bar magnets, one within the casing and the 
other without. The indicating needle is fastened to the outer magnet 
and revolves in harmony with the variation in height of the mercury 
column. The magnet controlling device is identical in principle with 
that described in connection with Fig. 554. 

The autographic instrument, Fig. 559, is one of the most successful 
recording devices on the market and is finding much favor with engi- 
neers. Its operation is shown diagrammatically in Fig. 560. Two 
cylindrical mercury cups R and R', constituting the legs of a U-tube 
mercury manometer, are pivoted on knife edge P and are connected 
to the nozzle plug by the flexible tubes A and A' '. The instrument 
may be placed at any point below- the level of the nozzle plug, one instru- 
ment being sufficient for a number of nozzles. When there is no flow 
the mercury in the two wells is at the same level and the system is in 
perfect balance. When there is a flow the velocity pressure causes the 






TESTING AND MEASURING APPARATUS 



789 



mercury to flow from R to R f and the latter is lowered, carrying with it 
the recording pen C. The vertical distance on the chart between and 



To Nozzle Plug 



Balancing Weight 




Fig. 559. G.-E. Recording Steam Meter. 

C is a measure of the weight of steam flowing. (This system of mercury 
wells was used as early as 1886 in a steam meter designed by Parenty.) 



< — 


A J 


L 


A' 


— > 


R 








R' 




s) 


<•> 




7 


V 








Fig. 560. Principles of the G.-E. Recording Steam Meter. 

Pressure variation is automatically compensated for by the sliding 
weight W, the position of which relative to the knife edge is effected by 
the Bourdon tube G; thus for increasing pressures the gauge tube tends 



790 



STEAM POWER PLANT ENGINEERING 



to straighten out and raise the weight, thereby decreasing the leverage 
of well R'. Adjustment for quality is made by shifting the position 
of sliding weight W by hand. 

In Europe the Gehre-Hallwachs meter, based upon the same principles 
as the G.-E. recorder, is fitted with an integrating device, thus enabling 
the total quantity of flow to be read directly. In the latter instrument 

the wells are cone shaped and pro- 
portioned so that the leverage in- 
creases directly with the weight of 
the steam flowing instead of increas- 
ing with the square of the weight 
as with the G.-E. cylindrical cups. 
This uniform motion of the cups 
makes it comparatively easy to add 
an integrating mechanism. Refer- 
ring to Fig. 561, R is a small friction 
wheel mounted on the pen arm a 
and connected to gears c and d by 
the small shaft m; P is a clock-driven 
disk in contact with the friction 
wheel R. As the pen arm moves the 
wheel R in and out from the center 
of disk P, the speed of the small friction wheel is decreased or increased 
accordingly. The revolutions of R are transmitted to the integrating 
mechanism e so that the total flow may be read directly from the 
dials. 

444. St. Johns Steam Meter. — In the groups of meters described 
above the indicating and recording mechanism is actuated by the natural 
velocity of the steam. In the St. Johns, Sargent, Gehre-Hallwachs, 
Storrer, Eckardt, and Venturi steam meters the velocity is increased by 
throttling and the pressure drop is utilized in actuating the mechanism. 
The weight of steam flowing through the orifice may be calculated from 
the following modification of equations (286) and (287) : 




Fig. 561. 



Counting Mechanism for 
Steam Meters. 



(290) 



W = AK V Pl - p 2 , 
in which 

W = pounds discharged per second. 
A = area of the orifice, square feet. 
K = coefficient determined by experiment and includes the density 

of the steam. 
Pi and P2 = pressure on the upper and lower side of the orifice, 
pounds per square inch. 



TESTING AND MEASURING APPARATUS 



791 



In some of the meters the pressure drop pi — p 2 is maintained con- 
stant and the variation in the area A actuates the indicating mechanism, 
and in others the area is made constant and the variation in pressure 
drop operates the mechanism. 

Fig. 562 represents a section through a St. 
Johns steam meter, illustrating the throttling 
type with a floating valve. This meter was 
placed on the market 20 years ago and still 
finds favor with many engineers. It records 
the weight of steam passing through the seat 
of an automatically lifting valve which rises 
and falls as the demand for steam increases 
or diminishes. 

Referring to the illustration, valve V is 
weighted so that a pressure in space A of 
2 pounds greater than in B is necessary to 
raise the valve off its seat. This pressure 
difference is constant for all positions of the 
valve. The plug is tapered so that the 
rise of the steam pressure is directly pro- 
portional to the volume of steam flowing through the seat. The 
movement of the valve is transmitted through suitable levers to an 
indicating dial and a recording pen so that the instantaneous and 
continuous rate of flow may be read at a glance. For a given pressure 




Fig. 562. St. Johns Steam 
Meter. 




DRAFT G»UGE 



%£ 



SIMPLE U TUBE 




Fig. 563. Different Forms of Manometer Pressure Gauges. 



and quality of steam, the indicating dial and chart may be calibrated 
to read the weight of discharge directly, corrections being made for 
variations in pressure and quality. The manufacturers guarantee the 
readings of the chart to be within 2 per cent of condenser measurements 
for a total pressure range of 10 pounds from the mean pressure at which 
the chart is calibrated. 




792 STEAM POWER PLANT ENGINEERING 

The chief drawback to this instrument is inherent to all meters of the 
direct type in that they are bulky and the steam line must be taken down 
for the installation. The total hourly flow may be obtained by inte- 
grating the curve. Tests of this meter made by the author were in 
accordance with the guarantee of the manufacturer for continuous flow 
and for moderate changes in the rate of flow. For rapid fluctuations in 
flow the results were not so satisfactory, the greater error lying in the 

difficulty of integrating the curve correctly. 
445. Pressure Gauges. — The Bourdon type 
of gauge, either autographic or indicating 
(Fig. 564), is the most familiar and satis- 
factory means of measuring pressures up 
to 1500 pounds per square inch or more, 
although diaphragm gauges are also used 
and both are employed as vacuum gauges. 
For the latter purpose, however, the mer- 
curial vacuum gauge has the advantage 
of greater accuracy and is not subject to 
Fig. 564. Bourdon Pressure derangement. Bourdon gauges should be 

frequently standardized by comparison with 
a gauge of known accuracy, a mercury column, or a gauge tester. 

For measuring very low pressures, such as are found in boiler flues 
or gas mains, indicating or recording diaphragm gauges may be had, 
but some form of U-tube manometer is generally employed, the design 
best adapted to the purpose depending upon the accuracy required. 
The simple U-tube (Fig. 563) when filled with mercury may be used for 
pressures limited only by the inconvenience due to length of tubes, or 
with water as the fluid, for pressures only a fraction of an ounce per 
square inch. Where greater accuracy is required than can be obtained 
with the simple U-tube, some modification may be employed, such as 
the Ellison draft gauge with one inclined leg which magnifies the reading 
several times. A form of sensitive gauge is sometimes used which 
depends upon the use of two fluids of different specific gravity, as oil 
and water. 

Pressure Gauges: Power, Aug. 15, 1911, p. 239; Aug. 18, 1908, p. 286. 

446. Measurement of Temperature. — For power-plant purposes mer- 
curial thermometers are most convenient for measuring temperatures 
up to 400 degrees F., and are inexpensive. For higher temperature, 
up to say 800 degrees F., they are also adapted, but must be made of 
special glass and the space above the mercury filled with nitrogen under 
pressure to prevent vaporization of the mercury. Such thermometers 



TESTING AND MEASURING APPARATUS 



793 



must be used intelligently and should be standardized from time to 
time, since they are subject to considerable change. The Bureau of 
Standards at Washington, D.C., is prepared to furnish certificates for 
which a nominal charge is made. 

Fig. 565 shows a form of thermometer which is much used where a 
continuous autographic record is required. It depends for its operation 
upon the pressure produced by a fluid, liquid or gaseous, contained in 
a small bulb and exposed to the temperature to be measured. The 
pressure is transmitted to the recording mechanism through a flexible 




Fig. 565. Bristol Recording Pyrometer. 



capillary tube which may be of considerable length. Such thermom- 
eters are suitable for feed water, flue gas, and temperatures not exceed- 
ing 1000 degrees F. 

Fig. 566 illustrates a form of electrical pyrometer employing thermo- 
couples which has come into wide use as a reliable means of measuring 
temperatures up to 2600 degrees F. The couples most frequently used 
are composed of platinum and platinum-rhodium, platinum and plati- 
num-iridium, copper and copper-constantan, and copper and nickel, 
the first named being adapted to the higher ranges of temperature. 
The electromotive force set up, when the thermo-j unction is heated, is 
proportional to the temperature and is measured by means of a sensitive 
milli voltmeter which is usually graduated to read temperature directly. 



794 



STEAM POWER PLANT ENGINEERING 



Thermo-couples may be made to give an autographic record by means 
of a thread recorder. 




Fig. 566. Bristol Thermo-Electric Pyrometer. 

Fig. 567 shows the element of an electrical thermometer based upon 
the change in resistance of a platinum wire when subjected to change 
in temperature. The resistance, in terms of temperature, is measured 
by a Whipple indicator, a convenient and portable form of Wheat- 
stone bridge, or may be autographically recorded by means of a Callen- 
dar recorder. Resistance thermometers of this type are very sensitive 
and accurate, not easily deranged, and are limited in range only by the 
fusing points of the platinum and the porcelain protecting sheath. 



PLATINUM COIL 




A PLATINUM WIRE LEADS 



Fig. 567. Element for Callendar Resistance Pyrometer. 

For higher temperatures and for obtaining the temperatures of in- 
closed spaces above about 900 degrees F., such as boiler furnaces, 
annealing ovens, and kilns, various forms of optical and radiation 
pyrometers have been devised. In such devices no part of the instru- 
ment is exposed to the temperature to be measured and hence suffers 
no injury from this cause. Optical pyrometers are based upon the 
measurement of the brightness of the hot body by comparison with a 



TESTING AND MEASURING APPARATUS 



795 



standard. The Wanner optical pyrometer is shown in Fig. 568. After 
standardizing by comparison with an amyl-acetate lamp, it is only 
necessary to focus the instrument upon the source of heat to be measured 
and the temperature is read on the graduated scale. 



FLAME 
GAUGE 



^AMYL-ACETAT 
LAMP 




Fig. 568. Wanner Optical Pyrometer in Position for Standardizing. 

Radiation pyrometers depend upon the measurement of the heat 
radiated from the hot body. The Fery radiation pyrometer, Fig. 569, 
is the best-known instrument of this type. When focused upon the 



TO GALVANOMETER 




Fig. 569. Fery Radiation Pyrometer. 



source of heat a cone of rays of definite angle is reflected by means of 
the mirror upon a thermo-couple located in its focus. The electromotive 
force set up is measured in terms of the temperature of the source of heat 
by a millivoltmeter. Neither the couple nor any part of the instrument 



796 



STEAM POWER PLANT ENGINEERING 



is ever subjected to a temperature much above 150 degrees F. The 
indications are practically independent of the distance from the source 
of heat, and the range is without limit. 

TABLE 127. 
TYPES OF THERMOMETERS IN GENERAL USE. 











Range in Degrees F. 




Principle of Operation. 


Type. 


for which they 










can be used. 


Expansion. . 




. . . Those depending on the 


Gas 


-400 to +2900 
— 35 to +950 






change in volume or 


Mercury, Jena glass, 






length of a body with 


and nitrogen. 








temperature. 


Glass and petrol ether. 

Unequal expansion of 
metal rods. 


-325 to +100 
to 950 


Transpiration and vis- Those denendine' on the 


The Uehling 


to 2900 


cosity. 




flow of gases through 
capillary tubes or small 
apertures. 






Thermo-electric . . . 


. . . Those depending on the 


Galvanometric 


— 400 to +2900 






electro-motive force 










developed by the dif- 










ference in temperature 










of two similar thermo- 










electric junctions op- 










posed to one another. 






Electric resistance. . 


. . .Those utilizing the in- 


Direct reading on indi- 


— 400 to +2200 






crease in electric resist- 


cator or bridge and 








ance of a wire with 


galvanometer. 








temperature. 






Radiation. . 




. .Those depending on the 


Thermo-couple in focus 
of mirror. 


300 to 4000 






heat radiated by hot 








bodies. 


Bolometer 


— 400 to Sun 


Optical .... 




. . Those utilizing the 


Photometric compari- 
son. 








change in the bright- 








ness or in the wave 


Incandescent filament 








length of the light 


in telescope. 


1100 to Sun 






emitted by an incan- 


Nicol with quartz plate 








descent body. 


and analyzer. 




Calorimetric . 




. .Those depending on the 


Platinum ball with 
water vessel. 


32 to 3000 






specific heat of "a body 








raised to a high tem- 










perature. 






Fusion 




. . Those depending on the 


Alloys of various fusi- 
bilities. (Seger cones.) 


32 to 3350 






unequal fusibility of 








various metals or 










earthenware blocks of 










varied composition. 







The Uehling pyrometer depends for its operation upon the flow of gas 
between two apertures, thus: Air is continuously drawn through two 
apertures by a constant suction produced by an aspirator. So long as 
the air has the same temperature in passing through these orifices there 



TESTING AND MEASURING APPARATUS 



797 



is no change in the partial vacuum in the chamber between them ; if, 
however, the air passing through the first opening has a higher tem- 
perature than that passing through the second, the vacuum in the 
chamber will increase in proportion to the difference in temperature 
since the volume of air varies directly with the temperature. In the 
application of this principle, the first aperture is located in a nickel 
tube which is exposed to the heat to be measured, while the second ap- 
erture is kept at a uniform lower temperature. This style of pyrom- 
eter is made to indicate and record and the indicating and recording 
mechanism can be placed at a distance from the main instrument. 

Table 127 embodies in outline the principles and temperature ranges 
of the various types of thermometers in use. Temperature ranges 
verified by U. S. Bureau of Standards. 

Modern Methods of Temperature Measurements: Cassier's Mag., June, 1909, p. 
99. High Temperature Measurements: Eng. and Min. Jour., Sept. 2, 1911, p. 447; 
Power, Aug. 2, 1910, p. 1376; Engineering, Feb. 9, 1912, Bui. No. 2, Bureau of 
Standards. 



y 'Maximum .Economy-.^ 

' for Given Conditiona \ 




WILSEY 

RELATIVE EFFICIENCY 

\ INDICATOR 

\ADKINS, YOUNG & ALLEN C» v ' 

\ CHICAGO U.S.A. / 

\No. a Typo-E / 






-4— !•- 



Leads to 
Current 



JJJ t tjH = 

Resistance Coil 

at Point where 

Gases leave Boile 



»j>>t i — 

Resistance Coil in 
Combustion Chamber 



Leads to Combustion Chamber 



Leads to Uptake 



Fig. 570. Wilsey Relative Efficiency Indicator. 



447. Wilsey Relative Efficiency Indicator. — Fig. 570 shows the 
general principles of the Wilsey relative efficiency indicator for indi- 
cating the relative boiler and furnace efficiency in any given installation. 
It consists essentially of two platinum resistance coils, a modified 



798 STEAM POWER PLANT ENGINEERING 

Wheatstone bridge, and a millivoltmeter. One of the coils is placed 
in the combustion chamber, the other in the uptake, and the indicating 
mechanism is mounted on the front of the boiler in plain view of the 
fireman. The ratio between the temperature of the combustion chamber 
and the uptake is a function of the boiler and furnace efficiency. The 
indicator is calibrated for each installation, so that when the boiler and 
furnace are operated at maximum efficiency under the given conditions, 
the indicating needle points to " maximum efficiency " or 100 per cent. 
If excess air is admitted the needle shows a lower relative efficiency 
due to the reduction in temperature, and if the air supply is deficient 
the needle swings beyond the predetermined " maximum efficiency." 
This device considers the efficiency of combustion solely from a quali- 
tative standpoint regardless of capacity. When installed in connection 
with a boiler flow meter the fireman is able to note the capacity and 
relative efficiency at a glance. Tests of this apparatus at the boiler 
plant of the Armour Institute of Technology gave reliable and con- 
sistent results. 

For a description of the Blonck " Boiler-Efficiency " meter see Elec. 
Rev. and Wes. Elec, Oct. 12, 1912; Prac. Engr. U. S., Feb. 15, 1913. 

448. Power Measurements. — Instruments for the measurement of power 
may be divided into two general classes, direct and indirect. The former 
involve the direct measurement of force and linear velocity or torque 
and angular velocity and the latter give the equivalent in other forms 
of energy. Direct power measuring appliances include the various 
speed indicators, transmission and absorption dynamometers, and the 
indirect include ammeters, voltmeters, watt-hour meters, boiler flow 
meters, and the like. In all power measurements the time or speed 
factor is readily determined but the force or torque factor, or equivalent, 
often involves considerable labor and the use of costly and complicated 
apparatus. The various conversion factors for the measurement of 
work, power, and duty are given in Appendix J. 

449. Measurement of Speed. — The following chart gives a classi- 
fication of a number of well-known instruments for determining linear 
and angular velocities. 

f Hand < Worm and Wheel. 



Counters 



Tachometer or 
Speed Indicators 



\ Gear Train. 
I Continuous | Electrical. 

(Centrifugal.. . j^ht, 



Electrical. 

Resonance \ Frahm's. 

ri,«v«« MH r,i, i Electro-magnetic 

Chronograph { Tuning Fo * k 






TESTING AND MEASURING APPARATUS 799 

The most commonly used device for speed determinations is the hand 
speed counter, consisting of a worm, worm wheel, and indicating dials. 
The errors to be corrected are principally those due to slipping of the 
point on the shaft, and to the slip of the gears in the counting device 
in putting in and out of operation. In some of the better grade of in- 
struments the gears are engaged or disengaged with the point in con- 
tact with the shaft. In the latter design a stop watch, actuated by the 
disengagement gear, minimizes the error likely to occur in hand manipu- 
lation. 

The continuous counter consists of a series of gears arranged to oper- 
ate a set of indicating dials. It may be operated by either rotary or 
reciprocating motion. The rate of rotation is calculated from the read- 
ings of the counter. 

All tachometers indicate directly the speed of the machine to which 
they are attached and are independent of time determination. ■ The 
most commonly used devices depend upon the centrifugal force of 
revolving weights for their operation. The indicating needle is at- 
tached to the weights in such a manner that the number of revolutions 
per minute is read directly from the position of the needle on the dial. 
These instruments should be calibrated for accurate work because of 
the number of wearing parts. 

Liquid tachometers consist essentially of small centrifugal pumps dis- 
charging into a vertical tube. The height of the indicating column is 
a function of the speed of rotation. 

Electrical tachometers are miniature dynamos, the voltage being a 
measure of the speed of rotation. These instruments are accurate and 
readily attached but necessitate the use of a delicate and costly volt- 
meter. The indicating mechanism may be placed at any distance 
from the small dynamo and in this respect has a marked advantage 
over the other types of speed indicators. 

The resonance tachometer affords a convenient method of measuring 
speeds over a wide range. It consists of a number of steel reeds of 
different periodicity mounted side by side on a suitable frame. When 
used to measure the speed of an engine or turbine the instrument is 
placed on or near the bed plates and the slight under or over balance 
causes the proper reed to vibrate in unison. 

450. Steam-engine Indicators. — This subject has been extensively 
treated by various authorities and a general discussion would be with- 
out purpose. For indicated horse power, testing indicator springs, and 
analysis or indicator diagrams see articles XII, XIV, and XX, "Rules 
for Conducting Steam Engine Tests," A.S.M.E., code of 1898. See also 
Preliminary Code for 1912, Jour. A.S.M.E., Nov., 1912. 



800 STEAM POWER PLANT ENGINEERING 

451. Dynamometers. — Dynamometers for measuring power are of 
two distinct types, absorption and transmission. In the former the 
power is absorbed or converted into energy of another form while in the 
latter the power is transmitted through the apparatus without loss, 
except for minor friction losses in the mechanism itself. 

The ordinary Prony brake is the most common form of absorption 
dynamometer. In the various forms of Prony brakes the power is 
absorbed by a friction brake applied to the rim of a pulley. For low 
rubbing speeds and comparatively small powers it affords a simple and 
inexpensive means of measuring the actual output. 

The Alden absorption dynamometer is a successful form of friction 
brake and has a wide field of application. It has been constructed in 
large sizes and is adapted to all practical ranges of speed. For a descrip- 
tion of rope brakes and the Alden absorption dynamometers see Ap-- 
pendix No. 19, p. 1848, Jour. A.S.M.E., Nov., 1912. 

Water brakes are finding much favor with engineers for high-speed 
service. There are two types, the Westinghouse and the Stumpf. In 
the former the rotor consists of a simple drum with serrated periphery 
revolving in a simple casing, the inner surface of which is serrated in 
a manner similar to the rotor. The resistance is produced by friction 
and impact, and the power is converted into heat which is carried away 
by the circulating water. The casing is free to turn about the shaft 
but is held against rotation by a lever arm. The torque of the lever 
arm is determined as in a Prony brake. A brake of this design, 2 feet 
in diameter and 10 inches wide will absorb about 3000 horse power at 
3500 r.p.m. In the Stumpf type the rotor consists of a number of 
smooth disks mounted side by side on a common shaft. The casing is 
divided into a number of compartments corresponding to the division 
of the rotor. There is no contact between rotor and casing. The 
friction between the disks and water and the water and casing tends 
to rotate the latter and the torque is measured in the usual way. In 
either type the power output is readily controlled by the water supply. 

Pump brakes and fan brakes are also used as absorption dynamometers. 
The latter are commonly used in connection with automobile engine 
testing. 

Electromagnetic brakes are occasionally used for power measurements. 
They consist essentially of a metal disk or wheel revolving in a mag- 
netic field. The resistance or drag tends to revolve the field casing 
and the torque is measured in the usual way. 

An electric generator mounted on knife edges forms the basis of the 
Sprague electric dynamometer. The prime mover drives the armature 
of the generator and the reaction between armature and field is counter- 



TESTING AND MEASURING APPARATUS 801 

balanced by suitable weights. The output is conveniently regulated 
by a water rheostat. 

Transmission dynamometers are seldom used for testing prime movers 
and are ordinarily limited to small power measurements. In some 
instances, however, as in marine service, transmission dynamometers 
afford the only practical means of approximating the net power de- 
livered to the propeller. For comparatively small power measurements 
may be mentioned the Morin, Kennerson, Durand, Lewis, Webber and 
Emerson transmission dynamometers, and for large powers, the Denny 
and Johnson electrical torsion meter and the Hopkinson optical torsion 
meter. For detailed descriptions of these appliances consult " Experi- 
mental Engineering, " Carpenter and Diederichs, Chap. X. 

452. Flue Gas Analysis. — It has been shown (paragraph 20) that 
the products of combustion, commonly called flue gases, resulting from 
the complete oxidation of coal with theoretical air supply consist chiefly 
of nitrogen and carbon dioxide, with lesser amounts of water vapor and 
sulphur dioxide. It was also shown that with a deficient air supply 
the flue gases may contain carbon monoxide and varying amounts of 
hydrocarbon. If excess air was used in the combustion of the fuel free 
oxygen would be present in the gases. Evidently an analysis of the 
flue gases offers a basis for judging the efficiency of combustion. The 
first step in the analysis and the most important one is the obtaining 
of a representative sample. Since the gases in the breeching and flues 
may be far from homogeneous great care must be exercised in getting 
a true average sample. (See Apparatus and Methods for Sampling and 
Analysis of Furnace Gases, U. S. Bureau of Mines, Bui. No. 12, 1911.) 

The analysis as ordinarily made in commercial practice is called 
volumetric, although in reality it is based upon the determination of 
partial pressures. According to Dalton's laws (paragraph 239) when 
a number of gases are confined in a given space each gas occupies the 
total volume at its own partial pressure, and the total pressure is the 
sum of all the partial pressures. When one of the gases is absorbed by 
a suitable medium and the remaining gases are compressed back to the 
original total pressure, a volume decrease is found, and if the tempera- 
ture remains constant this decrease represents the volume absorbed. 

The apparatus usually employed for volumetric analysis consists of 
a graduated measuring tube into which the gases are drawn and accu- 
rately measured under a given pressure, and a series of treating tubes, 
containing the necessary absorbing reagents, into which they are trans- 
ferred until absorption is complete. The Or sat apparatus, Fig. 571, 
forms the basis of nearly all of the portable appliances on the market 
for analyzing flue gases and the ordinary products of combustion. In 



802 



STEAM POWER PLANT ENGINEERING 



this apparatus a measured volume, representing an average sample of 
the gas, is forced successively through pipettes containing solutions of 
caustic potash, pyrogalic acid and cuprous chloride in hydrochloric 
acid, respectively, thus absorbing the carbon dioxide, the oxygen and 
the carbon monoxide, the contraction of volume being measured in each 
case. The apparatus as originally constructed is bulky and fragile 
and slow in its absorption of gas. 




Fig. 571. Standard Orsat Apparatus for Flue Gas Analysis. 

453. The Williams Improved Gas Apparatus is a marked improve- 
ment over the standard Orsat in that the objections cited above are 
obviated. In addition to the elimination of these objectionable features 
provision is made in the " Model A" type for the determination of 
illuminants, hydrogen and methane along with the three gases mentioned 
above. Referring to Fig. 572, A, B, C, and D are pipettes containing 
the necessary reagents for absorbing, respectively, C0 2 , illuminants, 2 , 
and CO. M is a graduated measuring flask connected at the bottom 
with water-level bottle W and at the top with the various pipettes. F 
is a hard rubber pump for taking gas sample directly from the source of 
supply, thereby eliminating the inaccuracy and annoyance of collection 
over water and transference. P is a portable case containing a spark 



TESTING AND MEASURING APPARATUS 



803 



coil and batteries for exploding the methane and hydrogen remaining 
in the burette after the other constituents have been removed. When 
extreme accuracy is desired mercury is used as the displacement medium 
in the leveling bottle since water absorbs C0 2 to a certain extent. For 
a complete description of this apparatus with sample calculations see 
paper read by F. M. Williams before Division of Industrial Chemists 
and Chemical Engineers, American Chemical Society, Washington, 
D. C, Dec. 28, 1911. 




Fig. 572. Williams Improved Gas Apparatus. 



454. "Little" Modified Orsat Apparatus. — Fig. 573 illustrates a 
modified Orsat apparatus as used by the Arthur D. Little Company of 
Boston, Mass. The right half of the device is the ordinary Orsat 
apparatus and the left portion constitutes the sampling attachment. 
The gases are drawn from the source of supply through rubber tube (2) 
into the sampling pipette (3) and out through rubber tube (1) to the 
aspirator. The latter may be operated by steam or water. When a 
sample is being collected the three-way cock on the glass header is closed 
and the mercury in the sampling tube (4) is allowed to drain through 
the movable overflow into the mercury retainer. The overflow is 
lowered at a constant rate by clockwork. Two driving pulleys afford 
seven different rates of movement downward of the overflow, thereby 
enabling a continuous sample to be collected at constant rate over any 
period from J to 24 hours. Instantaneous samples may be drawn off 
and analyzed as often as desired and with practically no delay to the 
continuous sample. For further details see Power, July 16, 1912, p. 77. 



804 



STEAM POWER PLANT ENGINEERING 



For many practical purposes it is sufficient to determine the carbon 
dioxide. A number of satisfactory appliances are on the market which 
give continuous autographic records of the percentage of C0 2 on clock- 
driven charts. These devices, however, are rather expensive and 
usually beyond the appropriation of small boiler plants. 



TOP VIEW OF THREE-WAY COCK 
^ENLARGED) 




Fig. 573. Modified Orsat Apparatus. — Arthur D. Little Co. 

455. Simmance-Abady C0 2 Recorder. — Fig. 574 illustrates the 
general principles of the Simmance-Abady C0 2 Recorder. The opera- 
tion is as follows: A continuous stream of water enters reservoir K 
through inlet X and overflow at 0. A portion of the stream flows into 
tank A through pipe F and causes bell float B to rise. As the float 
rises it permits bell D of the extractor to fall. When float B reaches 
the top of its stroke it raises valve stem E, trips the valve and causes 
the water to siphon out of tank A through siphon tube G. The lower- 
ing of the water level allows the bell to sink. As it falls it draws up 
the water-sealed extractor bell D and creates a partial vacuum under 
the latter. Flue gas then flows from the source of supply through P 
and H into the bell. The mass of water discharged from siphon tube 
G into the small vessel V beneath it overcomes the counterweight Q 
and closes the balance valve H, thereby entrapping a fixed volume of 
gas in the extractor bell. The stream of water which is continually 
flowing into tank A causes the float B to rise and the bell D to sink, as 






TESTING AND MEASURING APPARATUS 



805 



before. The lowering of bell D forces the entrapped flue gas through 
the caustic potash solution in vessel M into water-sealed recorder 
bell J. The displacement of bell J will be less than that of bell D by 
the volume of C0 2 absorbed in vessel M. The percentage of C0 2 in 
the flue gas is thus indicated by the position of the bell J with reference 
to the graduated scale N. The pen mechanism is attached to bell J 




To Boiler Room Indicator 
To Recording Gauge 



Caustic- Drip 



< -Absorption Chamber 

A 



Gas Inlet* 
fCaustic Overflow 

c^Y Fig. 575. Principles of the Uehling Gas 

Fig. 574. Simmance-Abady C0 2 Recorder. Composimeter. 

and records the percentage of C0 2 by the length of lines on a clock- 
driven chart. These samples are analyzed and the lines are drawn at 
three-minute intervals. The small water aspirator at X is for the pur- 
pose of exhausting gas continuously from the pipes connecting the 
recorder to the boiler, thereby insuring true samples at the time of 
absorption. Auxiliary pipe P is connected to main gas lead P. 

456. The Uehling Composimeter is another successful instrument for 
continuously recording the percentage of CO2 in the flue gas. The 
principles of this apparatus are illustrated in Fig. 575. The device 
consists primarily of a filter, absorption chamber, two orifices, A and 



806 



STEAM POWER PLANT ENGINEERING 



B, and a small steam aspirator. Gas is drawn from the usual source 
by means of the aspirator through a preliminary filter located at the 
boiler, and then through a second filter as illustrated in the diagram. 
From the latter the gas passes through orifice A, thence through the 
absorption chamber and orifice B to the aspirator where it is discharged. 
The C0 2 is absorbed by the caustic potash solution in the absorption 
chamber. This reduces the volume and causes a change in tension 
between the two orifices in proportion to the C0 2 content of the gas. 
This variation in tension is indicated by the water column, as shown, 
and is transmitted by suitable piping to the recording mechanism 
which may be placed at a considerable distance from the boiler room. 
457. Moisture in Steam. — Several forms of calorimeters are avail- 
able for determining the quality of steam. The simplest as well as 



-Thermome.tet 




Fig. 576. 



To Atmosphere- 

A Typical Throttling Calorimeter. 



the most satisfactory, if the percentage of entrained moisture is not 
beyond its range, is the throttling calorimeter, Fig. 576. In this device 
the sample of steam, which is taken from the steam pipe by means of 
the perforated nipple, is allowed to expand through a very small orifice 
into a chamber open to the atmosphere. The excess of heat liberated 
serves first to evaporate any moisture present and then to superheat 
the steam at the lower pressure. From the observed temperature and 
pressures it is easy to calculate, with the aid of steam tables, the per- 
centage of moisture in the original sample. 

The limit of the throttle calorimeter depends upon the steam pressure 
and is about 3 per cent of moisture at 80 pounds pressure and about 
5 per cent at 200 pounds. For steam containing greater percentages 
of moisture the separating calorimeter, Fig. 577, is sometimes used. 



TESTING AND MEASURING APPARATUS 



807 



This instrument is virtually a steam separator and mechanically sepa- 
rates the moisture from the sample of steam. The water thus separated 
collects in a reservoir provided with gauge glass and graduated scale, 
while the dry steam passes through an orifice to the atmosphere. The 
weight of dry steam per unit of time is indicated on the gauge, calcu- 
lated according to Napier's rule, or may be determined by condensing 
and weighing. The accuracy of the moisture determination is greatly 
affected by the difficulty of obtaining true samples of steam containing 
large percentages of moisture. 



ERMOMETER WELL 




DRAIN COCK 



Fig. 577. Carpenter Separating 
Calorimeter. 



Fig. 578. Ellison Universal Steam Calorimeter. 



Fig. 578 shows the Ellison universal steam calorimeter, which com- 
bines the superheating and throttling principles and is adapted to 
steam of any degree of wetness. The separating chamber is provided 
with a gauge glass, not shown, for indicating the weight of water which 
accumulates only when the steam is too wet to be superheated. 

Throttling Calorimeters: Power, Dec, 1907, p. 891; Trans. A.S.M.E., 17-151; 
175, 16-448; Engr. U. S., Feb. 15, 1907, p. 219. 

Separating Calorimeters: Trans. A.S.M.E., 17-608; Engr. U.S., Feb. 15, 1907, p. 219. 

Universal Calorimeter: Trans. A.S.M.E., 11-790. 

Thomas Electrical Calorimeter: Power, Nov., 1907, p. 791. 



808 



STEAM POWER PLANT ENGINEERING 



458. Fuel Calorimeters. — The analysis and heat evaluation of 
fuel require considerable time and skill and much costly apparatus, 
hence in most power plants it is customary to depend upon a specialist 
to whom samples are submitted from time to time. In many large 
stations, however, the conditions often warrant the establishment of a 
testing laboratory equipped for the proximate analysis of coal and the 
determination of the calorific value of the solid, liquid or gaseous fuel 
used. The Mahler bomb calorimeter illustrated in Fig. 579 is the 
most accurate and satisfactory device for solid and liquid fuels but is 
comparatively expensive. The instrument consists of a steel shell or 



INSULATION 
BOMB 

PLATINUM PAN 
WATER 

electrode 
ignition wire 
stirring device 
support for stirrer 
sensitive Thermometer 
oxygen tank 




Fig. 579. Mahler Bomb Calorimeter. 



"bomb" of great strength, lined with porcelain or platinum, into 
which a weighed sample of the fuel is introduced and burned on a 
platinum pan in the presence of oxygen under a pressure of about 300 
pounds per square inch. The charge is ignited by an electric current. 
During combustion the bomb is submerged in a known weight of 
water which is kept constantly agitated. The calorific value is calcu- 
lated from the observed rise in temperature due to the heat evolved, 
proper corrections being made for the water equivalent of bomb and 
appurtenances, heat given up by the igniting current, and for radiation 
or absorption of heat from the surrounding air. 

The Parr calorimeter, Fig. 580, is an inexpensive instrument, very 
simple in operation, and gives results which are sufficiently accurate 
for all practical purposes. The weighed sample of coal, together with 



TESTING AND MEASURING APPARATUS 



809 



a quantity of sodium peroxide which supplies the oxygen for com- 
bustion, is introduced into the cartridge. Means are provided for 
rotating the cartridge when submerged in the calorimeter, the at- 
tached vanes agitating the water to maintain uniform temperature. 



COMPRESSED 
FIBER 




Fig. 580. Parr Fuel Calorimeter. 



The charge is fired either electrically or by introducing a short piece of 
hot wire through the conical valve. The calorific value is calculated 
from the observed rise in temperature and the constants of the instru- 
ment. Among other forms of instruments, in more or less general use 
and which give very satisfactory results, may be mentioned the Car- 
penter, Thompson, Atwater and Emerson calorimeters. 

Comparison of Different Types of Calorimeters: Jour. Soc. Chem. Ind. (1903), 
22-1230. 



CHAPTER XVIII. 

FINANCE AND ECONOMICS. — COST OF POWER. 

459. General Records. — In all power plants, public or private, an 
itemized record of plant performance and cost of operation is of vital 
importance for the most economic results. In many states public 
utility corporations are required to submit an annual statement cover- 
ing the various details of operation, and in order to insure uniformity 
ruled and printed forms are furnished by the state. The private plant 
owner, on the other hand, is free to use his own judgment and may 
adopt any system of cost accounting or dispense with them entirely. 

The principal objects of keeping a system of records are (1) to enable 
the owner to compare the performance of his plant with current practice 
and to show him exactly what his plant is costing him, and (2) to enable 
the engineer to analyze the various records with a view of reducing all 
losses to a minimum. Power-plant records to be of value must be 
closely studied with a view to improvements. The mere accumula- 
tion of data to be filed away and never again referred to is a waste of 
time and money. 

Records should cover not only the daily, monthly, and yearly opera- 
tion of the plant but also, as permanent statistics, a complete analysis 
of each item of equipment. The value of such data cannot be over- 
estimated. The engineer will frequently find it greatly to his interest 
to have available at a moment's notice the complete details of his 
engines, boilers, generators, and other machinery, especially when it is 
required to renew a broken or worn-out part in case of emergency. 

The question of whether to purchase power or to generate it depends, 
chiefly, upon the relative cost of the two methods, although the absence 
of power-plant machinery and freedom from the coal and ash handling 
nuisance may be important factors. There is no doubt but that the 
central station can generate power cheaper than the small isolated 
plant, but in most cases it is a question not only of power, but also of 
supplying steam for heating and other purposes, and a careful study of 
all of the items entering into the problem is necessary for an intelligent 
choice. The service department of the large central station with its 
carefully maintained system of records has a strong advantage in pre- 
senting its arguments over the average private plant with its ill-kept 

810 



FINANCE AND ECONOMICS — COST OF POWER 



811 



and faulty system of accounting, and in some instances central-station 
service has been adopted simply because the engineer in charge was not 
in a position to prove positively that his own plant was the better 
investment. 

TABLE 128. 



PERMANENT STATISTICS. 

General Information. 



Date of installation 

Type of building 

Number of floors 

Number of offices 

Volume of building, cu. 

ft 

Type of heating system . 

Engine room, sq. ft 

Height of chimney, ft. . . 
Draft, inches of water . . 

Kind of grate or stoker . 

Kind of coal Ill 

Coal storage capacity, 
tons 

Capacity ice plant, tons 
in 24 hrs 

Capacity storage bat- 
tery, am. hrs 



Office 

18 

900 

10,860,000 

Webster 

6,840 

318 

3.5 

Jones 

Underfeed 

screenings 

450 

50 

None 



Total cost of building. . . 


$5,000,000 


Ground plan 


191X231 


Total office floor space, 




sq. ft 


400,000 


Height of building 


280 


No. of sides exposed. . . . 


3 


Radiator surface, sq. ft . 


100,000 


Boiler room, sq. ft 


5,400 


Number of elevators .... 


22 


Type of elevators | ^pressure 


Capacity of elevators, 




lb., each. . 


2,700 


Boiler pressure 


150 


Back pressure 


Atmospheric 


Part of bldg. lighted.... 


All 


Total cost of mechanical 




plant 


$650,000 





Engines. 


Generators. 


Motors. 


Boilers. 


Type 


Ball compd. 

5 

250 h.p. 


Crocker-Whpf»lf»r 




Number installed 


5 
150 kw. 


25 


5 


Rated capacity 


375 h.p. 




LIGHTS. 



Type 

Number installed 



Incandescent. 



Carbon 
25,000 



Tungsten 
5000 



Inclosed 
15 



A number of attempts have been made to standardize power-plant 
records but the results have been far from satisfactory because of the 
wide range in operating conditions. Each installation is a problem 
in itself and the items to be recorded must necessarily depend upon the 
size and character of the plant. A common mistake is to attempt too 
comprehensive a system with the result that after the novelty has 
ceased the labor of making the various entries becomes irksome, many 
of the items are omitted, guesses are substituted in place of actual 
observations, and the records are ultimately without value. A few 



812 



STEAM POWER PLANT ENGINEERING 



properly selected items, accurately recorded, are of vastly more impor- 
tance than an elaborate system of records indifferently maintained. 

460. Permanent Statistics. — Tables 128 to 131 are taken from the 
records of a large isolated station in Chicago and serve to illustrate the 
make-up of the " permanent statistics." The complete file covers each 
item of equipment and includes the various drawings, specifications, 
and guarantees for the entire mechanical equipment. Since these 
records do not vary with the operation of the plant they require no 
further attention, once they are compiled, except of course for such 
changes as may be made from time to time in the plant itself. 



TABLE 129. 

PERMANENT STATISTICS. 



Boilers. 



Make of boiler Stirling 

Total number in plant 5 

Date of installation .... 

Steam pressure, gauge 150 

Safety-valve pressure 160 

Type of safety valve Pop 

Area of grate, sq. ft .... 

Heating surface, sq. ft 3,500 

Superheating surface, sq. ft . None 

Number of steam drums ... 3 

Diameter of steam drums, in. 36 
Distance between steam 

drums, ft 3 

Thickness of shell, in f 

Thickness of head, in 

Diameter of steam nozzle, 



10 

2-4 in. 

2.5 

2 



Diameter of safety valve . . . 
Diameter of blow-off, in — 
Diameter of feed pipe, in. . . 
Temperature of flue, deg. 

Fah 450-490 

Temperature of feed water, 

deg. Fah 210 

Ratio of heating surface to 

grate area 41.6 

Kind of fuel 

Carterville, 111., Screenings 

Type of grate Green chain grate 

Rated horse power 375 



Number in battery 1 

Weight of boiler 62, 186 

Cost of boiler and fittings 

(each) $5,400 

Height of setting 17 ft. 9 in. 

Length of setting 17 ft. 4 in. 

Width of setting 15 ft. 3 in. 

Weight of setting 272,000 

Thickness of wall 

Side 20 in.; back, 15 in. 

No. of bricks, fire 6,590 

No. of bricks, common 19,600 

Dimensions of foundation 

15 ft. 2 in. X 17 ft. 4 in. 

Material of foundation 

Stone and concrete 

Cost of foundation and set- 
ting (each) $1,500 

Distance between batteries 4 ft. 6 in. 

Distance back of boiler. ... 17 ft. 6 in. 

Distance in front of boiler. . 16 ft. 6 in. 

Distance overhead 2 ft. 10 in. 



Number of tubes 

Diameter of tubes, in 

Length of tubes, ft 

Steam space, cu. ft 

Water space, cu. ft 

Kind of draft 

Inches of draft (maximum) 



337 

3.25 

12 to 14 

96 

643 

Forced 

3.5 



461. Operating Records. — The operating records of any plant bear 
the same relationship to the economical operation of that plant as the 
bookkeeping and cost accounting system bears to the manufacturing 
plant. The distribution of profit and loss in either case can only be 
obtained by itemizing the various factors involved and by grouping 
them in such a manner as to show at any time where improvement is 
possible. Commercial bookkeeping has been more or less standardized 






FINANCE AND ECONOMICS — COST OF POWER 



813 



and entails very little need of originality on the part of the bookkeeper, 
but the selection and maintenance of a system of power-plant records 
may require considerable study and experimenting, since each installa- 
tion is a problem in itself. The items included in the different forms 
depend upon the apparatus provided for weighing the coal and water, 



TABLE 130. 

PERMANENT STATISTICS. 

Main Engines. 



Make Ball Engine Co. 

Tandem or cross compound 

Cross Compound 

Number in plant 5 

Rated horse-power 250 

Average load 220 

Minimum load , 100 

Maximum load 315 

Best economy, lbs. per h.-p. 

hr 22 

Average economy, lbs. per 

h.-p. hr 25 

Steam pressure, gauge .... 150 

Receiver pressure, gauge . . 40 

Exhaust pressure Atmospheric 

Piston speed, ft. per min. . 604 

Type of governor 

Robb- Armstrong-Sweet 
Speed variation, per cent. . 1 



ft. 



Shop number . . . 
Height over all, 

Width, ft 

Length, ft 

Dimensions of foundation, 

ft 

Material of foundation. . . . 

Weight of engine, lbs 

Cost of engines and gener- 
ators 

Stroke, in 

Revolutions per min 

Weight of heaviest part, 
lbs 

R.p.m. of governor 

Diam. of flywheel, in 

Face of flywheel, in 

Weight of flywheel, lbs.. . . 



13 

14 

9 

13.5X10.5 
Concrete 

50,000 

$45,000 

16 

225 

28,000 

225 

72 

16.75 

4,000 



Diameter of cylinder, in 

Clearance, per cent 

Steam pipe diameter, in 

Exhaust pipe diameter, in 

Area of the port opening, in 

Diameter and length of main bearing, in. 
Diameter and length of crosshead pin, in, 
Diameter and length of crank pin, in. . . . 

Diameter and length of main shaft 

Diameter of piston rod 

Kind of piston packing 

Size of piston packing 

Kind of rod packing 

Area crosshead surface, sq. in 



H.P. 



14 

4.5 

5 

10 

10X1.75 

7X12 

4,3 

7 in., 3 

in., 14.5 ft. 

2.5 in. 

Snap ring 

■ in. square 

Metallic 

144.5 



L.P. 



22 
4.5 

7 

10 

14X2.5 

6X12 

4,3 

7 in., 3 

7 in., 14.5 ft, 

2.5 in. 

Snap ring 

f in square 

Metallic 

144.5 



Receiver. 





No. Heating Coils. 


Diameter, in 


7 

3.5 

1.05 


Length, ft 


Volume, cu. ft 


Diameter of drain pipe 





814 



STEAM POWER PLANT ENGINEERING 



the type and number of instruments available for measuring tempera- 
ture, pressure, and power, and the system adopted for keeping track of 
oil, waste, general supplies, and repairs. In large stations autographic 
recording and integrating appliances, which are to be found in nearly 

TABLE 131. 

PERMANENT STATISTICS. 



Feed Pumps. 



Date of installation 

Make Snow 

Number in plant 2 

Height, ft 3 

Length, ft 12 

Width, ft 4 

Weight of pump 5 tons 

Cost, each $965 

Steam pressure 150 

Back pressure \ 

Number of valves 32 

Character of valves 

Rubber, brass lined 
Area thro' valve seats, sq. in., 

per pump 12. 13 

Gallons of water per min., per 

pump 800 

Pounds of water per 24 hrs., 

average, actual 479,400 

Gallons of water per 24 hrs. . . 599.2 

Volume of air chamber, cu. ft. 3 

Shop number 24,572-3 



Diameter of steam cylinder. . 
Diameter of water cylinder . . 

Stroke 

Displacement per stroke, cu. 

ft 

No. of strokes per min., aver- 



16 
10 

12 

0.5454 



Diameter of suction 

Diameter of discharge 

Diameter of steam pipe 

Diameter of exhaust 

Diameter of steam drips 

Diameter of water drains .... 
Suction head, lb. per sq. in. . . 
Discharge head, lb. per sq. in. 
Kind of piston packing 

Outside packed plunger 

Size of piston packing 

Kind of rod packing Soft 

Size of rod packing f 

Temperature of feed water . . . 214 



12 
8 
5 
2.5 

4 

i 

2 

1 
2 

H 
175 



all strictly modern stations and represent but a small part of the first 
cost of the plant, greatly reduce the labor of keeping continuous records. 
In small plants the cost of autographic instruments may prove to be 
prohibitive and recourse must be had to the usual indicating devices. 
In the latter case, continuous records may be closely simulated by 
plotting the readings of the indicating appliances, say every 15 minutes, 
or even once every hour, and by connecting the points with a straight 
line. (See Figs. 581 to 583.) The oftener the readings are taken the 
smaller will be the error. Total quantities may be obtained by summing 
up the various items or by integrating the graphical chart by means 
of a planimeter. It is not sufficient to record monthly or yearly aver- 
ages. Daily and even hourly records are absolutely essential for 
maximum economy. The various losses may be reduced to a minimum 
only by an intelligent analysis of daily records. A number of forms 
taken from the files of various power plants are reproduced in this 
chapter under the proper subheadings and serve to illustrate current 
practice. 

Power Plant Records: Prac. Engr. U. S., July 1, 1912, p. 668; March 1, 1912, 
p. 242; Jan. 1, 1912, p. 36. Power, May 28, 1912,. p. 758. 



FINANCE AND ECONOMICS — COST OF POWER 815 

462. Output and Load Factor. — The output of a plant is usually 
stated in terms of the (1) average horse power, or equivalent, for a 
given period of time. (2) Unit output — horse-power hours, or equiva- 
lent. 

When the plant is operating at practically constant load it is suffi- 
ciently accurate for most purposes to express the output in horse power, 
or equivalent, per month or per year. When the output fluctuates as 
is the general case, it is best expressed in terms of unit output. For 
example, one horse power per year, 24 hours per day, and 365 days per 
year is equivalent to 365 X 24 = 8760 horse-power hours. If the full 
power is used throughout this time it matters little whether the charge 
is based on the flat rate (horse power per year) or the unit rate (horse- 
power hours); if, however, the power is used only half the time, the 
yearty cost per horse-power hour will be just double. 

The yearly load factor or simply load factor is the ratio of the actual 
yearly output to the rated yearly output measured on the twenty-four- 
hour basis. Thus: 

, - Yearly output, horse-power hours or equivalent /rh _., N 

Load factor = — ; F = = — J (291) 

Rated horse power, or equivalent X 8760 

The curve load factor or station load factor is the ratio of the yearly out- 
put to the rated output based upon the number of hours the plant is in 
actual operation. Thus, for an electric station: 

„ , . , Yearly output, kilowatt-hours ,™^ 

Curve load factor = ^ x , .; w * ' = — -^-, -. — . (292) 

Rated capacity X hours plant is in operation 

Much - confusion arises from the interpretation of the term "rated 
capacity." If rated below the maximum load it can sustain it is evident 
that a prime mover may operate with a load factor over 100 per cent, 
in which case the term is without purpose. The accepted definition of 
rated load in this connection is the maximum load which the prime 
mover can sustain continuously on a twenty-four-hour basis without 
overheating. Other definitions have been assigned to the term load 
factor and station factor, but the two stated above are more in accord 
with current practice. 

In any plant the great desideratum is a high load factor with greatest 
return on the investment. All the factors of expense included in the 
cost of power are then operating at maximum economy. High peak 
loads and low average loads necessitate large machines which are but 
little used and greatly increase the fixed charges. 

The demand factor is the ratio of the maximum demand to the con- 
nected load. There is a general tendency to overestimate the maxi- 
mum electric demand, due, in a measure, to the possibilities of all the 



816 



STEAM POWER PLANT ENGINEERING 



lights and motors being in use at one time. Practically speaking, such 
conditions are not likely to occur. Table 132 gives an idea of the value 
of the demand factor for various classes of service and may be used as 
a guide for determining the size of prime movers. 



TABLE 132. 

CENTRAL STATIONS, DEMAND FACTORS.* 
Demand factors compiled by Commonwealth Edison Company of Chicago. 

Class of Service. 



Lighting customers: 

Billboards, monuments, and department stores 

Offices 

Residences and barns 

Retail stores 

Wholesale stores 

Average 

Motor customers: 

Offices 

Public gathering places and hotels 

Residences and barns 

Retail stores 

Wholesale stores and shops 

Average 



Demand Factor. 



85.6 
72.4 
60.0 
66.3 
70.1 

59.8 



65.1 

28.7 
69.3 
61.2 
58.2 

59.4 



Demand factors compiled by Wisconsin State Commission from companies using Wright Demand Meter. 

Class of Service. 



Churches 

Clubs 

County and Federal buildings 

Depots 

Factories 

Hotels 

Laundries 

Livery stables 

Lodge and dance halls 

Offices 

Restaurants . . . , 

Saloons 

Schools 

Shops 

Shops — blacksmith 

Shops — machine 

Stores 

Theatres 



Demand Factor. 


Min. 


Max. 


56- 85 


28 




31- 


33 


75- 95 


53- 


56 


28 




60- 75 


52- 


58 


68 




57- 


87 


52- 


62 


62- 


92 


37- 


52 


55 




66 




37- 


54 


40-100 


49- 


89 



FINANCE AND ECONOMICS — COST OF POWER 



817 



TABLE 133. 

TYPICAL OPERATING CHART 

DAILY 
COAL TICKET. 



No. 



191 



6 a.m. to 6 p.m. 


6 p.m. to 6 a.m. 


Lbs. Coal. 


Lbs. Ashes. 


Lbs. Coal. 


Lbs. Ashes. 


































































































Kind of Coal. 


Kind of Coal. 






Fireman. 


Fireman. 



Coal used lbs. 

Ashes made lbs. 

Per cent ashes lbs. 

Water used gals. 

Water used lbs. 

Water lbs. to 1 lb. Coal. 

Boiler Water Test: Good. Med. Low. 



818 



STEAM POWER PLANT ENGINEERING 



TABLE 134. 
TYPICAL OPERATING CHART. 

DAILY POWER-HOUSE REPORT. 

The United Light and Power Co. 

Division 



.19.. 



Weather — 


Noon 














Engine No. 1 




started 

started 


M 

M 

M 

M 


stopped 

stopped 

off 

off 


M 

M 

M 

M 


Total time run . . 


Hr. 


Min. 


Engine No. 2 


Total time run .-. 























Noon AMPERE READINGS. 


12 00 


12 30 


1 00 


1 30 


2 00 


2 30 


3 00 


3 30 


4 00 


4 15 


4 30 


4 45 


5 00 


5 15 


5 30 


5 45 


6 00 


6 15 


6 30 


6 45 


7 00 


7 15 


7 30 


7 45 


8 00 


8 15 


8 30 


8 45 


9 00 


9 15 


9 30 


9 45 


10 00 


10 30 


11 00 


11 30 


12 00 


1 00 


2 00 


3 00 


4 00 


5 00 


5 15 


5 30 


5 45 


6 00 


6 15 


6 30 


6 45 


7 00 


7 15 


7 30 


7 45 


8 00 


9 00 


10 00 


11 00 



Coal used lbs. 

Cylinder oil pts. 

Engine oil pts. 

Waste lbs. 

Water cu. ft. 

Carbons 

Globes outer .... inner . . 



Coal Received on Track. 

Car No 

Initial 

Time placed m 

Time released m 

Weight lbs. 



Boilers in Service. 

No. 1 from m to m 

No. 2 from m to m 

No. 3 from m to m 

Washed No 

Blew No 



Ashes sold loads to . 



Material Received for Power House Use. 



Total Kilowatt Output. 
Read meter 12 o'clock noon 



Meter to-day Kw. 

Meter yesterday Kw. 

Diff 



Report here ANY interruption of service either arc or incandescent. 

Time off Cause 

Arc lights out 

Lights . . . 



Location Reported by 



FINANCE AND ECONOMICS — COST OF POWER 



819 



TABLE 135. 

TYPICAL OPERATING CHART. 

(Large Chicago Department Store.) 
Monthly Report. 





Average 
Outside 
Tempera- 
ture. 


Fuel. 


Supplies. 


Date. 


Coal. 


Ash. 


Oil Used, Gals. 


Waste 
Pounds. 


Total 




Kind. 


Pounds 
Burned. 


Cost 
Per 
Ton. 


jCost 
Per 
Day. 


Pounds 
Removed. 


Engine. 


Cylinder. 


Water to 
Building, 
Cu. Ft. 





Output. 






Engine-Hours Run. Boilers-Hours Run. 




Breeching. 


Boilers. 


Generators. 


1 


2 


3 


4 


5 


1 


2 


3 


4 


5 


6 


Draft. 




Pounds 

of 
Water 
Evapo- 
rated. 


Water 
Evapo- 
rated 
Per Lb. 
of Coal. 


Ampere- 
Hours. 


Kilo- 
watt- 
Hours. 


Tem- 
pera- 
ture. 



Heating System. 


Ventilating 

Plants, Hours 

Run. 


Refrigerating Plant. 


Repairs-Hours. 


Steam 
Pressure. 


Live 
Steam- 
Hours. 


Fan 
1 


Fan 
2 


Hours • T S* 
r>„_ Used, 
Run - Pounds. 


Ice 

Made, 
Pounds. 


Engine 
Room. 


Boiler 
Room. 


Miscel- 
laneous. 



In the original copy all of these items are conveniently grouped on one large form ruled for 31-day entries 
with space at bottom for total quantities and costs. In the reproduction only the headings are included. 



TABLE 136. 
TYPICAL OPERATING CHART. 

Annual Statement, First National Bank Building, Chicago. (Year 1911.) 
Expenditures. 



Coal $39,798.64 

Ash cartage 2,669.25 

Water 1,296.00 

Supplies and repairs: 

Elevators 1,948.08 



Engine room. 
Boiler room . . 
Supplies: 

Electrical 

Refrigerating . 
Steam fitting . 



631.80 
2,121.12 

1,952.01 

1,129.22 

296.63 

Heating 2,455.02 



Supplies: 

Oil waste and grease $1,728 .96 

Packing 500.00 

Plumbing 463.79 

Lamps 2,377.16 

Wages 26,986.84 

Petty cash 27.83 

Office 119.75 

Coal analysis 240.00 

Machine shop 183 . 42 



Total $87,891.24 



820 



STEAM POWER PLANT ENGINEERING 



TABLE 136 (Continued). 

Credits. 

Receipts from sale of current, steam power and the like. . $75,636.77 

Net cost of operation 1,225.47 

Assumed credits, account of elevator service, steam heat- 
ing, etc., for which payment is not made 70,260.39 

Net earnings including all credits . . . 58,005.92 



Unit costs — all 


expenses charged to power. 






• 




Cents. 


Pounds of coal, total for year. .... 

Total output, kilowatt-hours 

Coal per kilowatt-hour 


31,459,220 

1,854,400 

16.9 


Labor per kw.-hr 

Coal per kw.-hr 

Supplies per kw.-hr. . . 

Total 


1.45 
2.14 
1.13 




4.74 



463. Cost of Operation. — The cost of operation of power plants is 
conveniently divided into two parts: 

(1) Fixed charges. 

(a) Investment costs. 

(b) Administration costs. 

(2) Operating costs. 

464. Fixed Charges. — These cover all expenses which do not ex- 
pand and contract with the output. In very large plants they are 
usually divided into two parts: (a) the investment costs, which include 
interest, rental, depreciation, taxes, and insurance, and a reserve fund 
to cover depreciation of the investment, and (b) the administration costs, 
which include rental of offices, annual salaries of officers, and all other 
expenses not directly chargeable to the power plant. In the average 
plant the fixed charges comprise interest, rental, depreciation, taxes, 
insurance, and sometimes maintenance, though the latter is ordinarily 
included in the operating costs. 

In any system the total fixed charges per year are constant irre- 
spective of the load factor, since interest, taxes, depreciation, insurance, 
and maintenance go on whether the plant is in operation or not. The 
total fixed charges for a specific case are illustrated in Fig. 581 by a 
straight line. The cost per kilowatt-hour, however, decreases as the 
load factor increases. For example, with the plant operating con- 
tinuously at rated load (100 per cent load factor) the fixed charges 
per kilowatt-hour are 

65,000 



5000 X 8760 



= $0.00148. 



FINANCE AND ECONOMICS — COST OF POWER 

With 30 per cent load factor these charges are 
65,000 



821 



0.3 (5000 X 8760) 



= $0.00445 kilowatt-hour. 



The higher the load factor the greater is the amount of power pro- 
duced and the longer does the apparatus work at best efficiency. But 
the greater the power produced the larger will be the fuel consumption 
and the oil and supply requirements. The labor charges will be prac- 
tically constant. The total operating cost per year increases as the 
load factor increases, but not directly. (See Fig. 581.) The cost per 
-kilowatt-hour, however, decreases as the load factor increases. For 



(20000 




Yearly JLoad Factor— Per Cent 
Fig. 581. Influence of Load Factor on the Cost of Power at the Switchboard. (5000- 
kilowatt Electric Light and Power Station.) 

example, the operating costs per year with plant operating contin- 
uously at full load are $230,200. This gives 
230,200 



5000 X 8760 



1.00525 per kilowatt-hour. 



With 30 per cent load factor the yearly operating charges are $87,980, 
which gives 

87 Q80 
0.3 (5000 X 8760) = $ °-°° 67 per kilowa «- hour - 

Table 149 shows the influence of the load factor on the cost of power 
in two isolated stations of the same rated capacity, one operating 



822 



STEAM POWER PLANT ENGINEERING 



with the unusually high load factor of 80 per cent and the other operat- 
ing with the low load factor of 17 per cent. The former furnishes 
current for a large electro chemical concern in which the load is practi- 
cally constant. 

In general, the higher the load factor the greater becomes the ratio 
of the operating to the fixed charges, and extra investment may become 
advisable to secure the greatest economy possible. 

TABLE 137. 

AVERAGE INITIAL COST.* 

Steam Engine Power Plants. 

Simple Non-Condensing. 



Horse Power. 


Dollars per 
Horse Power. 


Horse Power. 


Dollars per 
Horse Power. 


10 


225.00 


60 


180.00 


• 20 


200.00 


70 


177.00 


30 


195.00 


80 


175.00 


40 


190.00 


90 


170.00 


50 


185.00 


100 


165.00 


Compound Condensing. 


100 


170.00 


700 


76.00 


200 


146.00 


800 


69.00 


300 


126.00 


900 


64.00 


400 


110.00 


1000 


60.00 


500 


96.00 


1500 


58.00 


600 


85.00 


2000 


55.00 


Triple Condensing. 


1000 


62.00 


4000 


52.00 


2000 


58.00 


5000 


50.00 


3000 


54.00 


6000 


48.00 



* Includes cost of buildings and entire equipment erected. 

On the other hand, when the load factor is low the fixed charges are 
the governing factor in the cost of power, and extra expenditures must 
be carefully considered, particularly if fuel is cheap. 

Fixed Costs in Industrial Power Plants: Engineering Digest, Apr., 1911, p. 293. 

465. Interest. — The rates of interest on borrowed money vary with 
the nature of the security. In the case of power plants the form of* 
security is usually a mortgage on the plant and equipment. If a 
builder has sufficient funds to construct the plant without borrowing, 
he should charge against the item "interest" the income which the sum 



FINANCE AND ECONOMICS — COST OF POWER 



823 



involved would bring if placed out at interest or if invested in his own 
business. In estimating the interest charges 5 per cent of the capital 
invested is ordinarily assumed unless specific figures are available. 

466. Depreciation. — This charge represents the gradual deterio- 
ration of a plant, resulting in its eventually wearing out. It is also 
assumed to represent the superannuation of a plant or the rate at 
which the apparatus is becoming obsolete. Thus, under the first 
assumption, if the useful life of an engine is 40 years, the rate of de- 
preciation, neglecting interest, is 2.5 per cent; if, however, it is assumed 



TABLE 138. 

COST OF MECHANICAL EQUIPMENT — ISOLATED STATIONS. 



Boilers (erected and set in masonry) : 

Horizontal-tubular 

Water-tube 

Steam engines : 

High-speed, simple direct-connected 

Medium-speed, compound non-condensing direct-connected. . . . 

Low-speed, compound condensing, belted 

Low-speed, simple, belted 

Gas engines 

Oil engines 

Gas producers 

Dynamos : 

Direct-connected to high-speed engine 

Belt-connected to engine 

Direct-connected to Corliss engine 

Switchboard 

Foundations 

Steamfltting — including auxiliary apparatus — such as feed heater. 

grease separator, exhaust head, tanks, covering, etc 



Per Kilowatt of 
Plant Capacity. 



$14-$18 
16- 20 

20- 25 
28- 35 
20- 25 
25- 30 
50- 60 
75- 85 

15- 20 

13- 16 
12- 15 

16- 20 
5- 10 
5- 10 

20- 30 



* By P. R. Moses before the A.I.E.E., Jan. 12, 1912. 

that the engine will become obsolete in 20 years and uneconomical for 
further operation, the rate of depreciation will be 5 per cent. It is 
difficult to assign a fixed rate of depreciation against any piece of ap- 
paratus, due to possible new developments which cannot be reckoned 
with in advance in computing the useful life of the apparatus. Again, 
depreciation cannot always be separated from current repairs and is a 
variable factor even in the parts of the same machine. It is, therefore, 
more or less of an approximation. The average life of various parts 
of a steam power plant is outlined in Table 140, but on account of the 
inability to assign fixed values to the useful life of any apparatus, and 
on account of the great number of appliances in even a small plant, it 
is customary to charge a fixed rate of depreciation against the entire 



824 



STEAM POWER PLANT ENGINEERING 



plant and thus avoid confusion and complexity. This very crude 
method usually results in overestimation in well-designed, well-operated 
plants and underestimation in poorly designed and badly managed 
installations. 

TABLE 139. 

COST OF MECHANICAL EQUIPMENT — STEAM TURBO-ELECTRIC GENERATING 

STATIONS.* 
2,000 to 20,000-kilowatt Capacity, Based on Maximum Continuous Capacity of Generators at 50° Rise. 



Dollars per Kilowatt. 



High. 



Low. 



Preparing site — Dismantling and removing structures from 
site, making construction roads, tracks, etc 

Yard Work — Intake and discharge flumes for condensing 
water, railway siding, grading, fencing sidewalks 

Foundations — Including foundations for building, stacks, and 
machinery, together with excavation, piling, waterproofing, 
etc 

Building — Including frame, walls, floors, roofs, windows and 
doors, coal bunker, etc., but exclusive of foundations, heat- 
ing, plumbing, and lighting 

Boiler-room Equipment — Including boilers, stokers, flues, 
stacks, feed pumps, feed-water heater, economizers, me- 
chanical draft, and all piping and pipe covering for entire 
station except condenser water piping 

Turbine-room Equipment — Including steam turbines and 
generators, condensers with condenser auxiliaries and water 
piping, oiling system, etc 

Electrical Switching Equipment — Including exciters of all 
kinds, masonry switch structure with all switchboards, 
switches, instruments, etc., and all wiring except for build- 
ing lighting 

Service Equipment — Such as cranes, lighting, heating, 
plumbing, fire protection, compressed air, furniture, per- 
manent tools, coal- and ash-handling machinery, etc 

Starting Up — Labor, fuel, and supplies for getting plant ready 
to carry useful load 

General Charges — Such as engineering, purchasing, super- 
vision, clerical work, construction, plant and supplies, 
watchmen, cleaning up 

Total cost of plant to owner, except land and interest during 
construction 



$0.25 


$.... 


2.50 


1.00 


6.00 


1.00 


12.00 


4.00 


24.00 


12.00 


22.00 


12.00 



5.00 

5.00 
1.00 

6.00 



$83.75 



2.00 

2.50 
.50 

3.00 



$38.00 



* By O. S. Lyford, J •., and R. W. Stoval, of Westinghouse, Church, Kerr & Company, before the 
Engineer's Society of Western Pennsylvania. 

In some of the very large stations depreciation is divided and charged 
as follows: 

(a) Complete depreciation due to wear and tear. This is charged to 
the proper maintenance account. 

(b) Obsolescence, inadequacy or destruction by any cause. This is 
written off from current income or from past accumulations of profit 
and loss account, or it is capitalized in the usual way. 



FINANCE AND ECONOMICS — COST OF POWER 825 

(c) Incomplete depreciation, due to wear and tear, likely to fall in 
large amounts at irregular intervals. This is provided for by a deprecia- 
tion fund. 

The depreciation fund is essentially a financial problem and the 
proper amount to be charged varies with every property, and should 
be carefully estimated either as a factor of the physical cost or of the 
gross earnings. 

TABLE 140. 

APPROXIMATE USEFUL LIFE OF VARIOUS PORTIONS OF STEAM POWER PLAN1 

EQUIPMENTS. 

Years. 

Buildings, brick or concrete 50 

Buildings, wooden or sheet iron 15 

Chimneys, brick 50 

Chimneys, self-sustaining steel 25 

Chimneys, guyed sheet-iron 10 

Boilers, water-tube 25 

Boilers, fire-tube 15 

Engines, slow-speed 25 

Engines, high-speed 15 

Turbines 25 

Generators, direct-current 25 

Generators, alternating-current 30 

Motors 20 

Pumps 25 

Condensers, jet 35 

Condensers, surface 20 

Heaters, open 30 

Heaters, closed 20 

Economizers 20 

Wiring 20 

Belts 7 

Coal conveyor, bucket 15 

Coal conveyor, belt 10 

Transformers, stationary 30 

Rotary converters 25 

Storage batteries 15 

Piping, ordinary., 12 

Piping, first class 20 

NOTE. — So much depends upon the design and the conditions of operation that no fixed 
values can be definitely assigned and the above figures should be used with caution. Practice 
shows that most power-plant appliances become obsolete long before the limit of their useful 
life is reached. 

The rate of depreciation in terms of interest and useful life is a simple 
problem in compound interest, and may be expressed: 

in which 

d = rate of depreciation, per cent of first cost. 

r = rate of interest. 

n = assumed life of the apparatus in years. 



826 



STEAM POWER PLANT ENGINEERING 



This is based on the assumption that the interest is compounded 
annually and that the apparatus is valueless at the end of n years. 
Table 141 has been calculated with this formula and gives the rate of 
depreciation for different rates of interest and different assumptions 
as to useful life of apparatus. 



TABLE 141. 



RATE OF DEPRECIATION. 















(Per Cent of First Cost.) 


* 












Rate of Interest, per Cent ? 




2 


2.5 


3 


8.5 


4 


4.5 


5 


5.5 


6 


7 


8 


9 


10 




2 


49.50 


49.37 


49.27 


49.14 


49.02 


48.90 


48.78 


48.66 


48.54 


48.31 


48.07 


47.84 


47.62 




3 


32.67 


32.51 


32.35 


32.19 


32.03 


31.87 


31.72 


31.56 


31.41 


31.10 


30.80 


30.51 


30.21 




4 


24.26 


24.08 


23.90 


23.72 


23.55 


23.39 


23.20 


23.03 


22.86 


22.52 


22.19 


21.84 


21.55 




5 


19.21 


19.02 


18.83 


18.65 


18.46 


18.28 


18.10 


17.91 


17.73 


17.40 


17.04 


16.73 


16.37 




6 


15.85 


15.65 


15.46 


15.26 


15.08 


14.89 


14.70 


14.52 


14.33 


13.97 


13.63 


13.29 


12.96 


a5 


7 


13.45 


13.25 


13.05 


12.85 


12.66 


12.46 


12.28 


12.09 


11.91 


11.15 


11.20 


10.87 


10.55 


1 


8 


11.65 


11.44 


11.24 


11.05 


10.85 


10.66 


10.47 


10.28 


10.10 


9.74 


9.40 


9.06 


8.74 


9 


10.25 


10.04 


9.84 


9.64 


9.45 


9.26 


9.07 


8.88 


8.70 


8.34 


8.00 


7.68 


7.36 


2* 
ft 


10 


9.13 


8.92 


8.72 


8.52 


8.33 


8.14 


7.95 


7.76 


7.58 


7.23 


6.90 


6.58 


6.27 


11 


8.21 


8.01 


7.80 


7.61 


7.41 


7.22 


7.04 


6.85 


6.68 


6.33 


6.00 


5.69 


5.40 


o 


12 


7.45 


7.25 


7.04 


6.85 


6.65 


6.46 


6.28 


6.10 


5.92 


5.60 


5.27 


4.97 


4.69 


SI 


13 


6.81 


6.60 


6.40 


6.20 


6.01 


5.83 


5.64 


5.47 


5.29 


4.96 


4.65 


4.36 


4.08 


3 


14 


6.26 


6.05 


5.85 


5.65 


5.46 


5.28 


5.10 


4.93 


4.75 


4.49 


4.13 


3.84 


3.58 


1 


15 


5.78 


5.57 


5.37 


5.18 


4.99 


4.81 


4.63 


4.46 


4.29 


3.97 


3.68 


3.40 


3.15 


P 


16 


5.36 


5.16 


4.96 


4.77 


4.58 


4.40 


4.22 


4.06 


3.89 


3.58 


3.30 


3.03 


2.78 


T3 


17 


4.99 


4.79 


4.59 


4.40 


4.22 


4.04 


3.87 


3.70 


3.54 


3.24 


2.96 


2.71 


2.47 


a 

1 


18 


4.67 


4.46 


4.27 


4.08 


3.90 


3.72 


3.55 


3.39 


3.23 


2.94 


2.66 


2.42 


2.19 


19 


4.37 


4.17 


3.98 


3.79 


3.61 


3.44 


3.27 


3.11 


2.96 


2.67 


2.47 


2.17 


1.95 


20 


4.11 


3.91 


3.72 


3.53 


3.36 


3.19 


3.02 


2.87 


2.71 


2.44 


2.18 


1.95 


1.95 




25 


3.12 


2.92 


2.74 


2.56 


2.40 


2.24 


2.09 


1.95 


1.82 


1.58 


1.36 


1.18 


1.75 




30 


2.46 


2.27 


2.10 


1.93 


1.78 


1.64 


1.50 


1.38 


1.26 


1.06 


0.88 


0.73 


0.61 




35 


2.00 


1.82 


1.65 


1.50 


1.36 


1.23 


1.10 


0.99 


0.89 


0.72 


0.58 


0.46 


0.37 




40 


1.65 


1.48 


1.32 


1.18 


1.05 


0.93 


0.83 


0.73 


0.64 


0.50 


0.38 


0.29 


0.22 




45 


1.39 


1.22 


1.07 


0.94 


0.82 


0.72 


0.62 


0.54 


0.47 


0.35 


0.26 


0.19 


0.14 




50 


1.18 


1.02 


0.88 


0.76 


0.65 


0.56 


0.42 


0.40 


0.34 


0.25 


0.17 


0.12 


0.09 



Example: A 1000-square-feet surface condenser and auxiliaries cost 
$3500. With interest at 5 per cent, required the rate of depreciation, 
assuming a life of 25 years. 

005 o no + 

= 2.09 per cent. 



d = 100 



(1 + 0.05) 25 - 1 



That is to say, if 2.09 per cent of the first cost is laid aside each year 
for 25 years at 5 per cent interest, compounded annually, it will equal 
the first cost at the end of this period. The sum thus laid aside is 
sometimes called the sinking fund. The solution is more readily 
obtained with the aid of Table 141; e.g., at the intersection of vertical 
column 5 (interest) and horizontal column 25 (life in years) we find 
the depreciation 2.09 per cent. This sinking-fund method of rating 



FINANCE AND ECONOMICS — COST OF POWER 



827 



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828 STEAM POWER PLANT ENGINEERING 

the depreciation is peculiarly adapted to power-plant practice, inas- 
much as a sum set aside at comparatively low rates of interest and 
compounded increases very slowly at first but more and more rapidly 
from year to year. This is precisely what happens in the deterioration 
of the plant. The loss of value is slight at first when the materials 
are new and usefulness is at a maximum, while towards the end of 
life both value and usefulness decline very rapidly. 

If the apparatus still has some value at the end of n years and if b is 
the ratio of the value of the old material to that of the new, the rate 
of depreciation becomes 

<*° i M ( i+ 1 r 7»-i ; < 294 > 

Example: In the foregoing problem, required the rate of depreciation 
if the value of the condenser outfit is $350 at the end of 25 years. 

Here b = 3500 = °- L 

100 X 0.05 (1 - 0.1) 
(1 + .05) 25 - 1 
= 1.97 per cent. 

That is, 1.97 per cent of $3500, or $68.95, laid aside each year for 25 
years at 5 per cent interest and compounded annually will equal $3500 — 
350, or $3150, at the end of this period. 

It is not supposed that an owner will regularly lay aside this sum 
annually, or take the trouble to arrange for its investment at current 
rates in the market or savings bank, since the money is probably worth 
more to him in his own business. In practice it is retained in his 
business or investments and is earning the rate of interest obtainable 
therein, but in determining the net profit or loss this depreciation item 
is nevertheless accounted for just as if it were actually placed in out- 
side investments. 

In appraising the present value of any apparatus in terms of the rate 
of interest and useful life the expression becomes 

r-™- vV&-\ ' (295) 

in which 

V = total depreciation, per cent of original cost. 
m = number of years apparatus has been in operation. 
n — assumed life of apparatus. 
r = rate of interest. 



FINANCE AND ECONOMICS — COST OF POWER 829 

Example: In the preceding example, required the present valuation 
of the condenser, assuming that it has been in use 15 years. 

m = 15, n = 25, r = 0.05. 
Substituting these values in (295), 

T7 inA (l + 0.05) 15 - 1 . K1 

V = 1Q0 (i + 0.05)25 _ ! = 45 -! P er cen *. 

That is, the condenser has depreciated 45.1 per cent of its original value 
and consequently is worth $3500 - 0.451 X 3500 = $1921.50. * 

Table 141 may be conveniently used in this connection. At the inter- 
section of vertical column 5 and horizontal columns 15 and 25 we find 
463. and 2.09 respectively. Dividing 2.09 by 4.63 we get 0.451 = 45.1 
per cent, the total depreciation. Depreciation is often taken care of 
under the different items pertaining to maintenance, and whenever a 
change or repair is necessary it is charged directly into expense as 
maintenance. 

Though usually considered separately, interest and depreciation are 
sometimes considered as a single item, and in this case the rate of de- 
preciation represents the sum which must be laid aside each year for 
the eventual renewal of apparatus plus the interest on the investment. 

Depreciation: Jour. Elec. Power & Gas, Feb. 25, 1911, p. 176; Elec. Rev. & Wes. 
Elecn., Apr. 23, 1910, p. 852, Dec. 2, 1911, p. 1126; Elec. Wld., Dec. 2, 1909, p. 
1349; Power, June 18, 1912, p. 878, Feb. 25, 1911, p. 176, Nov. 22, 1910, p. 2087, 
Aug. 9, 1910, p. 1439, July 5, 1910, p. 1227. 

467. Maintenance. — Maintenance usually refers to the expense of 
keeping the plant in running order over and above the cost of attend- 
ance. It includes cost of upkeep, replacement, and precautionary 
measures. This latter item includes the renewal of working parts, 
painting of perishable or exposed material, and replacing worn-out and 
defective material. Many engineers make no allowance for mainte- 
nance in the fixed charges and include these costs under supplies, 
attendance, or repairs. In a general way, when maintenance is in- 
cluded under the fixed charges, an annual charge of 2 per cent is con- 
sidered a liberal allowance, since most of the repair work comes under 
attendance. In street-railway practice maintenance is divided among 
the several parts of the system as follows: Buildings, steam appli- 
ances, electrical equipment, and miscellaneous. In this connection the 
maintenance becomes a part of the operating charges, since the various 
items vary widely from month to month. 

468. Taxes and Insurance. — Taxes vary from a fraction of one per 
cent to one and one-half per cent, depending upon the location of the 
plant. An average figure is one per cent of the actual value of the 



830 STEAM POWER PLANT ENGINEERING 

investment. Buildings and machinery are ordinarily insured against 
fire loss and boilers against accidental explosions, and accident policies 
are sometimes carried on all operating machinery. A fair charge for 
this item is one-half per cent. 

469. Operating Costs. — Operating costs are conveniently divided as 
follows : 

(1) Labor or attendance. 

(2) Fuel and water. 

(3) Oil, waste, and supplies. 

(4) Repairs and maintenance. 

In large stations it is often desirable to keep the expenses of the 
various departments separate from those of the power-plant proper. 
Thus in central stations the distributing system is an important branch 
and the attending expenses form a considerable portion of the total. 
They are therefore kept separate, since they are not strictly chargeable 
to power generation. In isolated stations the wages of elevator men, 
though in a general way a part of the power-plant expenses, are not 
included in the "labor and attendance" charge of the plant. Lamps 
are a large item of expense in a tall office building, and for this reason 
are often kept separate from supplies. 

470. Lsbor, Attendance, Wages. - — The minimum number of men 
required to handle a given plant is approximately a fixed quantity and 
it is seldom possible to so arrange the work that any material reduction 
can be effected. Until very recently it has been the universal custom 
to pay wages on a "flat rate" basis, that is, the attendant is given a 
fixed sum per day or month irrespective of the amount of work required 
or the economy of operation. In many cases, however, the bonus 
system has been successfully adopted. For example, in the boiler room 
the coal consumption is determined for a given period of time with 
ordinary careful firing, and the fireman is offered a reasonable per- 
centage on the saving of coal which he is able to effect over this record 
by special care and attention to the keeping of fires always in the best 
condition, avoiding the blowing off of steam, using as little coal as 
needed for banking fires, and in other ways. Where careful records are 
kept of supplies, repairs, and renewals, the bonus is also applicable to 
electricians, oilers, and other employees. 

Labor should always be estimated or recorded as so many dollars 
per month or per year and not merely in terms of the output unless 
the load factor is definitely known, otherwise comparisons are mis- 
leading. For example, consider two plants of 500 kilowatts capacity, 
each with labor charges, say, of $400 per month. Suppose the output 



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FINANCE AND ECONOMICS — COST OF POWER 833 

of one is 100,000 kilowatt-hours per month and that of the other 40,000 
kilowatt-hours per month. The monthly charges are evidently the 
same, viz., $400, but the cost per kilowatt-hour differs widely, being 
0.4 cent in the first case and 1 cent in the latter. 

The cost of labor varies so much with the location of the plant and 
the conditions of operation that general figures are of little value except 
as a rough guide. The figures in Tables 143 and 144 give average 
results for general practice. Specific figures will be found in Tables 
145 to 160. 

For a summary of labor costs in large central stations see " Central-Station Labor 
Costs," Electrical World, Nov. 16, 1912, p. 1031. 

471. Cost of Fuel. — Tables 145 to 160 give examples of the cost of 
fuel in different types and sizes of steam power plants. It will be 
noted that this item varies considerably even with plants of the same 
general class. So much depends upon the market price of the fuel 
itself that the item "cost per horse power or kilowatt-hour" gives 
little information concerning the economy of operation unless the 
price of the fuel, its calorific value, and the water rate of the prime 
movers are specified. In a general sense the cost of fuel will range 
from 40 to 70 per cent of the total operating expenses. In estimating 
the cost of fuel for a proposed installation due consideration should be 
given to the coal burned in banking fires, heat lost in blowing off boilers, 
and reduced efficiency in operating at underloads and overloads. For 
example, individual tests of a number of boilers in a large central 
station in Chicago gave an average evaporation of 6.1 pounds of water 
per pound of coal, actual conditions, whereas the evaporation deter- 
mined by dividing the total water fed into the boiler per year by the 
total consumption of coal gave only 5.2 pounds. The difference be- 
tween the two depends largely upon the standby losses in banking fires 
and in shutting down and starting up boilers. The following data 
from a number of installations in Chicago give an idea of the extent of 
these losses when burning Illinois coal : 

1. Five 400-horse-power boilers with Murphy furnaces required 
1300 pounds of coal for banking over night and 3800 pounds for firing 
up when cold. 

2. Five 500-horse-power boilers with chain grates required 3900 
pounds for banking 6 hours and 5000 pounds for banking over night 
and being put in service. 

3. One 400-horse-power boiler with Dutch oven required 1500 pounds 
for 7-hour bank and 2500 pounds for 13-hour bank. 

4. One 66 by 18 return tubular boiler, hand fired, required 500 pounds 
for 12-hour bank. 



834 



STEAM POWER PLANT ENGINEERING 



Current practice gives an average efficiency (based on yearly opera- 
tion) of boiler and furnace of 70 per cent for pumping stations running 
at practically full load, 65 per cent for large lighting and power stations 
with yearly load factor of 0.50 or more, and 60 per cent for similar 
stations with load factor between 0.35 and 0.45. For very low load 
factors, 0.25 and under (as in connection with large manufacturing 
plants, tall office building, and other plants operating on a 12-hour basis), 

this efficiency sel- 
dom exceeds 50 per 
cent. With these 
figures as a guide 
the cost of fuel per 
unit output may 
be roughly approxi- 
mated. 

In Europe the 
"locomobile" type 
of steam power 
plant has attained 
an extremely high 
degree of heat effi- 
ciency as will be seen 
from the curves in 
Fig. 582. The most 
economical result 
shown, namely 0.87 
pound of coal per 
developed horse- 
power hour, is 
equalled only by 

1900 1910 i , 

Fig. 582. Development of the Steam Power Plant. 0Ur DeSt g as "P r °- 

(Locomobile Type.) ducer plants. 

472. Oil, Waste, and Supplies. — These items approximate from 2 to 
10 per cent of the total operating expenses. Tables 144 to 160 give some 
idea of current practice in different classes of power plants. 

473. Repairs and Maintenance. — This item ordinarily refers to the 
cost of keeping the plant in running order over and above the cost of 
labor or attendance, and depends upon the age and condition of the 
plant and the efficiency of the employees. Tables 144 to 160 give the 
cost of repairs and maintenance for a wide range in power-plant practice. 

474. Cost of Power. — The actual cost of producing power depends 
upon the geographical location of the plant, the size of apparatus, the 




FINANCE AND ECONOMICS — COST OF POWER 



835 



TABLE 145. 

COST OF ONE HORSE POWER PER YEAR, SIMPLE ENGINES. NON-CONDENSING,, 

10-HOUR BASIS, 308 DAYS PER YEAR. 

(Wm. O. Webber, Engineering Magazine, July, 1908, p. 563.) 



Size of plant horse power 

Cost of plant per horse power. 

Fixed charges at 14 per cent 

Coal per horse-power hour, in pounds. . . . 

Cost at $4.00 per ton 

Attendance, 10-hour basis 

Oil, waste, and supplies 

With coal at $5.00 per ton 

With coal at $4.50 per ton 

With coal at $4.00 per ton 

With coal at $3.50 per ton 

With coal at $3.00 per ton 

With coal at $2.50 per ton 

With coal at $2.00 per ton 



20 

$200.00 

28.00 

12.00 

66.00 

30.00 

6.00 

146.50 

138.25 

130.00 

121.75 

113.50 

105.25 

97.00 



40 

$190.00 

26.60 

10.00 

55.00 

20.00 

4.00 

119.35 

112.47 

105.60 

98.72 

91.85 

84.97 

78.10 



60 
$180.00 
25.20 

9.00 
49.50 
15.00 

3.00 
105.07 
98.80 
92.70 
86.51 
80.32 
74.13 
67.95 



80 
$175.00 
24.50 

8.00 
44.00 
13.00 

2.60 
95.10 
89.60 
84.1® 
78.60 
73.10 
67.60 
62.10 



TABLE 145a. 

COST OF ONE HORSE POWER PER YEAR, COMPOUND CONDENSING ENGINES, 

10-HOUR BASIS, 308 DAYS PER YEAR. 

(Wm. O. Webber, Engineering Magazine, July, 1908, p. 564.) 



Size of plant horse power 

Cost of plant per horse power. . . . 
Fixed charges at 14 per cent. 
Coal per horse-power hour, pounds 

Cost of fuel at $4.00 per ton 

Attendance, 10-hour basis 

Oil, waste, supplies 

Total 

With coal at $5.00 per ton 

With coal at $4.50 per ton 

With coal at $4.00 per ton 

With coal at $3.50 per ton 

With coal at $3.00 per ton 

With coal at $2.50 per ton 

With coal at $2.00 per ton 



100 


200 


300 


400 


500 


$170.00 


$146.00 


$126.00 


$110.00 


$96.00 


23.80 


24.40 


17.65 


15.40 


13.45 


7.0 


6.5 


6.0 


5.5 


5.0 


38.50 


35.70 


33.00 


32.00 


27.50 


12.00 


10.00 


8.60 


7.25 


6.20 


2.40 


2.00 


1.72 


1.45 


1.24 


76.70 


68.10 


60.97 


56.10 


48.39 


86.40 


77.10 


69.22 


61.90 


55.29 


81.50 


72.60 


65.07 


58.10 


51.79 


76.70 


68.10 


60.97 


56.10 


48.39 


71.90 


63.70 


56.82 


50.50 


45.04 


67.00 


59.20 


51.67 


46.70 


41.49 


62.30 


54.75 


48.59 


43.00 


38.83 


57.45 


50.25 


44.47 


40.10 


34.64 



600 

$85.00 

11.90 

4.5 

24.70 

5.40 

1.08 

43.08 

49.28 

46.18 

43.08 

3G.98 

36.88 

33.83 

30.73 



Size of plant horse power 

Cost of plant per horse power. . . . 

Fixed charges at 14 per cent 

Coal per horse-power hour, pounds 

Cost of fuel at $4.00 per ton 

Attendance, 10-hour basis 

Oil, waste, supplies 

Total 

With coal at $5.00 per ton 

With coal at $4.50 per ton 

With coal at $4.00 per ton 

With coal at $3.50 per ton 

With coal at $3.00 per ton 

With coal at $2.50 per ton 

With coal at $2.00 per ton 



700 


800 


900 


1000 


1500 


$76.00 


$69.00 


$64.00 


$60.00 


$58.00 


10.65 


9.65 


8.95 


8.40 


8.12 


4.0 


3.5 


3.0 


2.5 


2.0 


22.00 


19.20 


16.50 


13.75 


11.00 


4.70 


4.15 


3.75 


3.50 


3.25 


0.94 


0.83 


0.75 


0.70 


0.65 


38.29 


33.83 


29.95 


26.35 


23.02 


43.79 


39.73 


34.05 


29.80 


25.77 


41.04 


36.28 


32.00 


28.05 


24.39 


38.29 


33.83 


29.95 


26.35 


23.02 


35.54 


31.48 


27.87 


24.60 


21.64 


32.79 


29.03 


25.80 


22.90 


20.27 


38.04 


27.18 


23.75 


21.20 


18.89 


27.29 


24.23 


21.70 


19.47 


17.52 



2000 

$56.00 

7.85 

1.5 

8.25 

3.00 

0.60 

19.70 

21.75 

20.72 

19.70 

18.67 

17.65 

16.60 

15.57 



836 



STEAM POWER PLANT ENGINEERING 



TABLE 146. 

SHOWING CAPITAL COSTS OF PLANTS INSTALLED AND ANNUAL COSTS OF POWER 
PER BRAKE HORSE POWER, AVERAGE WORKING CONDITIONS. 

(H.Von Schon, Engineering Magazine, May, 1907.) 

Class I. — Engines: Simple, Slide-Valve, Non-Condensing. 
Boilers: Return Tubular. 



Size of Plant, 
Horse Power. 


Engines, 

Boilers, etc., 

Installed. 


Capital Cost of Plant per 
Horse Power Installed. 


Annual Cost 

of 10 Hours 

Power per 

B.H.P. 


Annual Cost 

of 20 Hours 

Power per 




Building. 


Total. 


B.H.P. 


10 


$66.00 
56.00 
48.70 
44.75 
43.00 


$40.00 
37.00 
35.00 
33.50 
31.00 


$106.00 
93.00 
83.70 

78.25 
74.00 


$91.16 
76.31 
66.46 
59.49 
53.95 


$180 76 


"20 

30 


151.48 
131 68 


•40 


117 74 


£0 


106 46 







Class II. — Engines: Simple, Corliss, Non-Condensing. 
Boilers: Return Tubular. 



30 


$70.70 
62.85 
59.00 
56.00 
50.00 
44.60 


$35.00 
33.50 
31.00 
30.00 
27.50 
25.00 


$105.70 
96.35 
90.00 
86.70 
77.50 
69.60 


$61.14 
55.50 
50.70 
47.42 
43.86 
40.55 


$117.70 


40 


107.10 


50 


97.73 


60 


91.34 


80 


85.41 


100 


79.19 







Class III. — Engines: Compound, Corliss, Condensing: 
Boilers: Return Tubular with Reserve Capacity. 



100 


$63.40 
53.70 
50.10 
45.90 
43.55 
41.25 
40.50 
39.00 


$28.00 
24.00 
20.00 
18.00 
16.00 
14.00 
13.00 
12.00 


$91.40 
77.70 
70.10 
63.90 
59.55 
55.25 
53.50 
51.00 


$33.18 
29.83 
28.14 
26.27 
24.84 
23.73 
23.56 
23.26 


$60.05 


150 


54.63 


200 

300 


51.72 
48.83 


400 


46.12 


500 


44.21 


750 


44.02 


1000 


43.71 







FINANCE AND ECONOMICS — COST OF POWER 



837 



design, conditions of loading, system of distribution, and the method 
of accounting. Tables 145 to 160 compiled from various sources give 
the detailed costs of a large number of central and isolated stations. 
Table 147 and Figs. 586, 587, and 588 give the fundamental relations 
between the various items entering into the cost of power for various 
types of plants of over 30,000 kilowatts rated capacity. These data 
are taken from a paper presented by H. G. Stott at a meeting of the 
Toronto section of the American Institute of Electrical Engineers, 
Toronto, Ont., December, 1908. The figures have been brought up to 
date (June, 1912) by Mr. Stott and show what is actually being done 
to-day in large plants of the size stated above. 

TABLE 147. 

DISTRIBUTION OF MAINTENANCE AND OPERATION COSTS IN POWER PLANTS 
HAVING A MAXIMUM LOAD OVER 30,000 KILOWATTS. 

_ (H. G. Stott.) 



Maintenance. 

1. Engine room, mechanical 

2. Boiler or producer room 

3. Coal- and ash-handling apparatus . 

4. Electrical apparatus 

Operation. 

5. Coal 

6. Water 

7. Engine room, labor 

8. Boiler or producer room, labor . . . 

9. Coal- and ash-handling, labor 

10. Ash removal 

11. Electrical labor 

12. Engine room, lubrication 

13. Engine room, waste, etc 

14. Boiler room, lubrication, etc 

Relative operating cost, per cent 

Relative investment, per cent 

Probable average cost per kilowatt . . . 
Probable fixed charges 



Recip- 


Steam 


Recip- 
rocating 
Engines 


Gas 


Gas 

En- 


rocating 


Tur- 


and Low- 


En- 


and 
Steam 
Tur- 
bines. 


Steam 


bine 


pressure 


gine 


Plant. 


Plant. 


Steam 
Tur- 
bines. 


Plant. 


2.59 


0.51 


1.55 


5.18 


2.84 


4.65 


4.33 


3.55 


1.16 


1.97 


0.58 


0.54 


0.44 


0.29 


0.29 


1.13 


1.13 


1.13 


1.13 


1.13 


61.70 


55.53 


46.48 


26.52 


25.97 


7.20 


0.65 


0.61 


3.60 


2.16 


6.75 


1.36 


4.06 


6.76 


4.06 


7.20 


6.74 


5.50 


1.81 


3.05 


2.28 


2.13 


1.75 


1.14 


1.14 


1.07 


0.95 


0.81 


0.54 


0.54 


2.54 


2.54 


2.54 


2.54 


2.54 


1.78 


0.35 


1.02 


1.80 


1.07 


0.30 


0.30 


0.30 


0.30 


0.30 


0.17 


0.17 


0.17 


0.17 


0.17 


100.00 


77.23 


69.91 


52.94 


47.23 


100.00 


75.00 


80.00 


110.00 


96.20 


125.00 


93.75 


100.00 


137.50 


120.00 


11% 


11% 


11% 


12% 


11.5% 



Hy- 
drau- 
lic. 



0.51 



1.13 



1.36 



2.54 
0.20 
0.20 



5.94 
100.00 
125.00 
11% 



For steam-turbine plants larger than 60,000 kw. the cost per kilowatt may be 
reduced to $75.00. 



838 



STEAM POWER PLANT ENGINEERING 



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STEAM POWER PLANT ENGINEERING 



TABLE 149. 

COST OF POWER. 
Examples of Isolated Station Practice. 





Large Office 
Building. 


Small Office 
Building. 


Manufactur- 
ing Plant, 
Electro- 
plating. 


Rated capacity kilowatts 

Yearly capacity kilowatt hours 

Actual load kilowatt hours 


500 

4,380,000 

670,000 

17 

17 


50 
438,000 
40,820 
9.1 
24.7 


500 

4,380,000 

3,500,000 

80 

80 


Yearly load factor per cent 

Curve load factor per cent 



Operating Charges, per Year. 



Labor 

Coal and ashes 


$6,050.00 

6,342.00 

642.00 

168.00 

395.00 

69.00 


$1,400.00 

960.00 

75.00 

90.00 

41.00 

182.00 


$12,300.00 
9,100.00 


Water 


Oil and waste 


210 00 


Lamps 

Repairs and renewals 


50.00 
1,008.00 


Total 


$13,666.00 


$2,748.00 


$22,668.00 



Fixed Charges, per Year. 



Interest (5 per cent) 


$3,500.00 
4,200.00 

350.00 
1,050.00 

900.00 


$325.00 

628.00 

30.00 

90.00 


$4,500.00 

5,400.00 

450.00 


Depreciation (6 per cent) 


Insurance (& per cent) 


Taxes (1A per cent) 


1,350.00 


Rental 








Total 


$10,000.00 


$1,073.00 


$11,700.00 







Cost per Kilowatt Hour, Cents. 



Operating charges 
Fixed charges 
Total cost 




.65 
.33 
.98 



FINANCE AND ECONOMICS — COST OF POWER 



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FINANCE AND ECONOMICS — COST OF POWER 



845 



TABLE 152. 

YEARLY OPERATING COSTS IN FOUR TYPICAL CENTRAL STATIONS, 

STATE OF MASSACHUSETTS 

Year ending June, 1909. 



Type of Prime Mover 

Rated station capacity, kw.. 
Output, millions of kw. hrs.. . 
Yearly load factor, per cent. . 
Total station operating force 
Cost of fuel, dollars per ton . . 
Coal per kw. hr 



Salem Elec- 
tric Light 
Co. 



6 Engines 

2500 

3.106 
14.2 
14 

4.51 

3.3 



Fitchburg Gas 

& Electric 

Co. 



3 Engines 

2000 

4.006 
22.9 
12 

4.52 

3.28 



Haverhill 
Electric Co. 



2 Turbines 
1 Engine 

2300 

3.721 
18.5 
13 

3.97 

3.27 



Maiden 
Electric Co. 



1 Turbine 
3 Engines 



4.715 



14 

3.78 
3.02 



Operating 


Costs, Cents 


per Kilowatt Hour. 




Coal 


0.740 
0.025 
0.027 
0.410 
0.034 
0.158 
0.011 


0.740 
0.015 
0.025 
0.308 
0.017 
0.041 
0.072 
0.024 


0.650 
0.019 
0.003 
0.285 
0.063 
0.073 
0.019 
0.040 


0.565 


Oil and waste 


020 


Water 


0.045 


Wages 


320 


Station building repairs 

Steam equipment repairs. . . . 
Electrical equipment repairs. 
Miscellaneous 


0.023 
0.072 
0.014 
021 








Total.... 


1.412 


1.242 


1.152 


1.08 



TABLE 153. 

COST OF POWER, CENTS PER KW. HOUR. STEAM-ELECTRIC CENTRAL STATIONS. 

Year ending June 30, 1908. 



Fuel 

Oil and waste 

Water 

Wages 

Station repairs 

Steam repairs 

Electrical repairs 

Miscellaneous 

Total 

Cost of fuel per ton 

Output, millions kilowatt 
hours per year 

Capacity of station, thou- 
sands of H.P 



Bos- 
ton. 



.462 
.008 
.024 
.192 
.015 
.042 
.056 
.023 



.822 
3.99 

88.5 

73.5 



Worcester. 



.703 
.027 
.034 
.360 
.012 
,055 
055 
000 



1.246 
4.79 

5.4 

5.90 



Lowell 



.710 
.009 
.008 
.262 
.020 
.020 
.009 
.022 



1.060 
4.75 

9.4 

7.39 



Fall 
River. 



.880 
.032 
.012 
.538 
.012 
037 
029 
080 



1.620 
4.68 

4.0 

4.43 



Mai- 
den. 



.635 
.017 
.032 
.342 
.035 
.072 
.014 
.033 



180 
49 



4.6 

4.87 



Cam- 
bridge. 



.690 
.019 
.055 
.347 
.021 
.059 
.046 
.000 



1.237 
4.40 

6.0 

6.75 



Lynn. 



.618 
.012 
.040 
.296 
.052 
.147 
.045 
.000 



1.210 
3.60 

8.7 

8.2 



846 



STEAM POWER PLANT ENGINEERING 



TABLE 154. 

COST OF ONE HORSE POWER PER YEAR IN STREET-RAILROAD SERVICE FOR 
DIFFERENT CLASSES OF ENGINES. 

1000-Horse-Power Plant. (R. C. Carpenter.) 

(Sibley, Journal of Engineering, November, 1904, p. 92.) 



Non-Condensing Engines. 

Simple slide-valve, aver- 
age 

Simple slide-valve, best. . . 

Simple Corliss, average 

Simple Corliss, best 

*Compound slide-valve. . . . 
Condensing Engines. 

Compound slide-valve, 
average 

Compound slide-valve, 
best 

Compound Corliss, average 

Compound Corliss, best. . . 



^n 


Tons per Horse 
Power per Year. 


oo o 
^h © 


o 












03 ¥? 

F - 


■a 


erest, 5 per Cen 
epreciation, 10 
mt. 


03 




B 


rs o 


8 

L 


O 

M 

CO 

7 


o 
W 

1 


3l Cost per 
ours per D 
r Ton. 


03 
I 


8 


e3 

Q 




03 
ft 


3 W & 


O 


|P5 


4.63 


10.14 


15.21 


20.28 


30.42 


7.30 


4.44 


4.60 


10.07 


15.10 


20.14 


30.20 


7.30 


4.44 


3.45 


7.55 


11.33 


15.10 


22.66 


7.30 


4.75 


3.01 


6.59 


9.89 


13.18 


19.78 


7.30 


4.75 


4.17 


9.05 


13.57 


18.10 


27.14 


6.90 


4.80 


3.25 


7.12 


10.68 


14.24 


21.36 


6.70 


4.72 


2.40 


5.25 


7.88 


10.51 


15.76 


6.50 


4.72 


2.36 


5.17 


7.74 


10.33 


15.48 


6.50 


5.28 


1.80 


3.94 


5.91 


7.88 


11.82 


6.10 


5.28 



■a! 

* o 
p US 



42.16 
41.94 
34.71 
31.83 
38.84 



32.78 

26.98 
27.26 
23-20 



* The compound slide-valve engine, running non-condensing, made, in this series of tests, a 
poorer record than the single Corliss. This may have been due to the extremely bad conditions 
of loading. It is, I think, a fact that this class of engine has not been a marked success for 
street railway wolk. 



TABLE 155. 

OPERATING EXPENSES OF BOSTON ELEVATED RAILWAY COMPANY. 

Average of All Stations at Switchboard, Oct. 25, 1912. 





Year. 




1905 


1906 


1907 


1908 


1909 


1910 


1911 


1912 


Rated capacity, kilowatts 


35,544 
43.7 
0.45 
0.17 


38,469 
43.0 
0.47 
0.17 


39,969 
45.0 
0.55 
0.18 


50,425 
37.0 
0.56 
0.21 


50,063 
37.0 
0.45 
0.20 


51,163 
41.5 

0.48 
0.17 


51,463 
41.9 
0.44 
0.17 


61,350 
36.4 


* Coal, cents per kilowatt-hour. . . 
Labor, cents per kilowatt-hour. . . . 


0.41 
0.17 


Supplies, cents per kilowatt-hour. . 


0.57 


0.60 


0.76 


0.86 


0.61 


0.58 


0.55 


0.52 


Total per kilowatt-hour 

Ratio operating expense to gross 


0.72 

0.68 
$3.1354 


0.77 

0.687 
$3.1859 


0.94 

0.69 
$3,572 


1.07 

0.67 
$3,568 


0.81 

0.654 
$3,209 


0.75 

0.642 
$3,283 


0.72 

0.646 
$3,346 


0.69 


Price of coal per ton 


$3,202 



* Coal included in supplies. 



t Repairs included in supplies and labor. 



FINANCE AND ECONOMICS — COST OF POWER 



847 



60 






















l. = 

2.= 


Recipr 
Steam- 


ocating 
turbine 


steam-plant, 
plant. 










/<T 


50 

DC 

< 

UJ 

> 

cc 

UJ 

°-40 
H 
H 
< 

O 

_l 

*30 

UJ 

Q. 
CO 

< 

_J 

§20 


3.= 


Recipr 


ocating 
turbi 


engine 
ne plant 


and low 


-pressui 


■e 








4.= 

5.= 

6— 


Gas-en 

Gas-en 

-Hydra 


gine pl£ 
gine an< 
xlic-plai 


nt. 

1 steam- 


turbine 


plant. 






















^\ 




















^ 












S 


^ 














^ 


^ 
















^ 


^ 
















3# 
















6n 


10 

































































20 



40 60 

PER CENT LOAD-FACTOR 



80 



100 



Fig. 583. Cost of Power in Large Power Plants with Maximum Load over 
30,000 Kilowatts. Coal at $3.00 per Ton. 14,500 B.T.U. per Pound. 



848 



STEAM POWER PLANT ENGINEERING 



60 






















1. = Reciprocating steam-plant. 

2.— Steam-turbine plant. 

3.— Reciprocating-engine and low-pressure 

turbine plant. 
4. = Gas-engine plant. 
5 .= Gas-engine and steam-turbine plant. 
6 . = Hydraulic-plant 












1 y 


50 






















10 








































*+ 


3 ^ 


30 




















5^ 




















""• 




20 












































3^ 
^2 




















10 













































20 



40 60 

Per Cent Load-Factor 



Fig. 584. Cost of Power in Large Power Plants with Maximum Load over 30,000 
kilowatts. Coal at $1.50 per Ton. 11,000 B.T.U. per Pound. 



FINANCE AND ECONOMICS — COST OF POWER 849 

TABLE 156. 

COST OF POWER. 

Pacific Gas and Electric Company. 

Kilowatt-hours generated by steam 85,707,854 

Kilowatt-hours generated by transmission 7,787,959 

93,495,813 
Kilowatt-hours sold 68,797,090 

Kilowatt-hours lost in distribution 24,698,723 

Per cent loss, 26.5. 

TOTAL COSTS. 

Revenue from sales $2,730,248. 00 

Cost of generation $729,315. 00 

Cost of distribution 347,182. 00 

Cost of administration 943,363.00 2,019,860.00 

Net earnings $710,388 . 00 

UNIT COSTS, CENTS PER KILOWATT-HOUR. 

Generation: Distribution: 

Labor ; 0.225 Labor 0.216 

Materials 0.731 Materials 0.098 

Repairs 0.104 Repairs . 191 

1 . 060 0.505 

Administration: Summary of Unit Costs: 

Labor . 271 Generation 1 . 060 

Materials 0. 082 Distribution 0. 505 

Legal Expenses . 021 Administration . 576 

Fire Insurance 0. 005 Interest 0. 006 

Bad Debts 0. 026 Depreciation 0.789 

Advertising . 008 2 936 

Damages to persons . 005 

Rental.. 0.005 

Taxes 0.153 

0.576 

TABLE 157. 

COST OF POWER. 

Year Ending June 30, 1911, Central-station Production at Fall River. 

Equipment: One 500-, two 1000-kilowatt turbo-alternators. 
Six 350-horse-power boilers. 

Total output: 5,764,466 kilowatt-hours. 

Load factor: 28 per cent. 

Total Operating Costs: 

Coal, 7726 tons at $3.67 $28,332.00 

Oil and waste 633 . 00 

Water 747.00 

Miscellaneous supplies 598 . 00 

Wages, 16 men 13,329.00 

Repairs, building 5,537 . 00 

Repairs, steam equipment 1,555 . 00 

Repairs, electrical 109 . 00 

$51,840.00 
Unit Operating Costs, Cents Per Kilowatt-hour: 

Fuel 0.49 

Wages 0. 23 

Supplies and Repairs 0.18 

0T90 



850 



STEAM POWER PLANT ENGINEERING 



16 














1 1 1 1 

FIXED CHARGES 

1 1 1 i 














1. 


Reciprocating steam plant. Cost $125.00 
-Steam-turbinc-plantr— eJst-$93r75-per-Kw 
» $75.00 \ « 


jerKw. 






14 






2, 














3. 


Eng 


ne and low pressure 


turbine plant. 










12 






4. 


Cost $100.00 per Kw. 
Gas engine plant. 


:ost $137.50 per Kw. 














5. 


Gas 


engine anc[ steam turbine 


plant. 










10 


CO 
HI 


\\\ 


6. 


Cost" 
Hyd 


$i20.( 

rauli( 


)U per km. 

i plant. Cost $125.00 


per Kw. 










cc 
< 


llV 








1 1 1 
TOTAL OPERATING 

i 


CHARGE 


:s 






8 


o 

Q 


l\ 




1- 


5 A 

3 


s above, co 


al@ 

-@.$1 


$3.00- 
50=- 


-1450( 
1000- 


B.t.u. p< 


;rlb. 
lb 






LU 
>< 
Ll 
























6 






























































4 
































^ 






























2 










Ssl> 


^<3" 


U<6 




























">- 



















































CO 
UJ 
CD 










































6" 


2 


< 






















jj-5' 








o 




-^ — 


HS 


EH^ 


^"-j 









-^ 


4' 


^ 




-t£ 2 ' 












r 


\A 






\ 


z 
r- 

— <- 


— 7^ 


„^- 


— 


— 











.___ 








,«3 


"¥ 






UJ 


s 














—^ 
















O 
i 


/> 














S 
































6 


< 

r- 

o 
































r- 































20 



40 



60 80 100 

PERCENT LOAD 



120 



140 



Fig. 585. 



Cost of Power in Large Power Plant with Maximum Load over 
30,000 kilowatts. 



FINANCE AND ECONOMICS — COST OF POWER 851 

TABLE 158. 

COST OF POWER AT THE POWER PLANT OF FITCHBURG YARN COMPANY FOR THE 
YEAR ENDING MARCH 25, 1911. 

Hewes & Phillips cross-compound Corliss engine, 26 and 56 by 60 inches : 

Boiler pressure per square inch 175 pounds. 

Receiver pressure per square inch 19 pounds. 

Hewes & Phillips independent-driven Venturi condenser. 

5 Dillon's Manning type boilers, 74-inch waist, 17-foot tubes: 

Grate area 192.40 square feet. 

Heating surface 10,449 square feet. 

Superheating surface 4662 square feet. 

LB. Davis gleaner exhaust heater: 

Water from 44 to 118 degrees 

LB. Davis gleaner heater: 

Exhaust of condensing engine and condensation from re- 
ceiver, water from 118 to 152 degrees 

Green fuel economizer, 4488 square feet heating surface: 

Water from 152 to 232 degrees 

Gases entering economizer 430 degrees 

Gases entering chimney 230 degrees 

Height of chimney 165 feet 

William A. Jepson coal, semi-bituminous: 

Cost of coal per ton « $4 . 50 

Maximum indicated horse power 1291 . 8 

Minimum indicated horse power 2107 . 8 

Average indicated horse power 1744 . 28 

COST OF POWER. 
4229.57 pounds coal per horse power (including banking and 

heating) $8. 46 

Labor 2.92 

Supplies and repairs 1. 11 

Total operating expenses $12 . 49 

Depreciation and interest $4. 01 

Taxes 0. 72 

Insurance 0.04 

Total fixed charges 4.77 

Total gross cost $17.26 

Deduct cost of heating . 58 

Total cost one horse power per year $16 . 68 

TABLE 159. 

COST OF OPERATION. 

Power Plant of the Hyde Park Electric Light Company, Boston, Mass. (1910). 

For the year ending June 30, 1910, total kilowatt-hours, sold 4,103,384 

Of the total 88.5 per cent was for street railway service or 3,636,390 

Total output at station, kilowatt-hours 4,357,648 

Total generator capacity in kilowatts 1,775 

Coal, bituminous at $3.94 and No. 3 Buckwheat at $3.17. 

NET COST AT SWITCHBOARD. 

Fuel $34,228. 04 equals $0.0078 kw.-hour 

Oil and waste 746. 45 

Water 707. 63 

Station wages 9,609. 76 equals $0.0022 kw.-hour 

Repairs, station building 607 . 44 

Repairs, steam equipment 2,966 . 14 

Repairs, electric equipment 652.71 

Minor tools 776.19 

Total $50,294. 36 equals $0.0115 kw.-hour 



852 



STEAM POWER PLANT ENGINEERING 



TABLE 160. 

CENTRAL-STATION STATISTICS. (STATE OF IOWA, 

Cities 1,000 to 2,000 Population. 



1909.) 



Station number. 



Population, thousands 
Consumers per 100 population 
Station Capacity: 

Kilowatts 

Watts per capita 

Ratio transformers to 

Load per kw. sta. cap.: 

Lamps, kw 

Motors, kw 

Heat and misc., kw 

Total connected, kw 

Yearly load factor 

Investment, dollars: 

Per kw. capacity 

Per capita 

Gross annual income, dollars: 

Per kw. capacity 

Per consumer 

Per capita 

Ratio expense to gross in- 
come 



1 


2 


3 


4 


5 


6 


7 


8 


9 


10 


1.1 

12 


1.2 
10 


1.4 
9 


1.5 

9 


1.5 

7 


1.5 
12 


1.8 
10 


2.0 
11 


2.0 
11 


2.0 


45 
41 
1.2 


50 
42 


48 
29 


100 
67 


60 
40 
1.0 


60 
40 


75 
42 
0.1 


120 
60 
1.7 


53 

27 
1.1 


60 
30 


0.7 
0.0 
0.0 
0.7 
32 

$244 
10 


2.4 
0.1 
0.0 
2.5 
22 

$370 
15 


'46**' 

$250 
9 


1.5 
0.2 
0.1 
1.8 
21 . 

$200 
13 




1.3 

0.2 


0.6 
0.0 


1.7 

0.0 


1.1 

0.0 


0.8 
0.0 




1.5 


0.6 
50 

$300 
13 


1.7 
47 

$208 
13 


1.1 


0.8 


$192 
8 


$200 
8 


$260 

7 


$133 
4 


$81 
26 
3.30 


$93 
39 
3.90 


$114 
43 . 
3.90 


$90 
67 
6.00 




$88 
29 
3.50 


$103 
42 
4.30 


$70 
37 
4.20 


$111 
27 
2.90 


25 "' 


80% 


75% 


65% 




72% 


76% 


83% 


36% 


81% 


64% 



Av. 
1-10 

1.6 
10 

67 
42 
1.0 

1.3 

.006 
.03 
1.3 
36 

$236 
10 

$94 
37 
4.00 

70% 





Cities 2,000 to 3,000 Population. 












11 

2.2 
12 

75 

34 
0.7 

2.3 
0.4 
0.1 

2.8 


12 

2.5 
6 

80 
32 


13 

2.5 

130 
52 


14 

2.7 
11 

110 
41 
1.2 

1.6 
0.5 
0.4 
2.5 

24 

$218 
9 

$94 
36 
3.80 
79% 


15 

2.8 
12 

135 
48 
0.9 

2.2 
0.0 
0.5 

2.7 
24 

$200 
9 

$67 
38 
3.20 

72% 


16 

3.0 
9 

175 
58 
1.2 

1.3 
0.0 
0.0 
1.3 
23 

$200 
11 

$101 
65 

5.90 
63% 


17 

3.0 
10 

150 

50 

D.C. 

1.2 
0.0 
0.0 
1.2 

$167 

8 

$67 
33 
3.30 


18 

3.0 
10 

120 
40 
1.3 

1.8 
0.8 
0.3 
2.9 
25 

$367 
15 

$132 
50 
3.30 
32% 


Av.' 




11-18 
2.71 


Consumers per 100 population 

Station capacity: 


10 
122 




44 




1.1 


Load per kw. sta. cap.: 


1.3 

0.0 

...... 

45 

$187 
6 

$69 
40 

2.20 
73% 


2.1 
0.0 
0.0 
2.1 

$138 

7 

'*44'" 

"72%' 


1.7 




0.2 


Heat and miscellaneous, kw 


0.2 
2.1 




28 


Invesin.ent, dollars: 


$292 
10 

$137 

40 
4.60 

62% 


$221 




9 


Gross income, dollars: 


$95 




43 




3.80 i 


Ratio expense to gross income. . . 


63% | 



Cities 4,000 to 53,000 Population. 



Station number 19 

Population, thousands 

Consumers per 100 population 
Station capacity: 

Kilowatts 

Watts per capita 

Ratio trans, to sta. cap 

Load per kw. sta. cap.: 

Lamps, kw 

Motors, kw 

Heat and misc., kw 

Total connected, kw 

Yearly load factor 

Investment, dollars: 

Per kw. capacity 

Per capita 

Gross income, dollars: 

Per kw. capacity 

Per consumer 

Per capita 

Ratio expense to gross in- 
come 



19 


20 


21 


Av. 


22 


23 


Av. 


24 


25 


26 


4.0 
11 


5.0 
11 


8.0 
9 


5.67 
10 


15.0 

7 


16.0 
3 


15.5 
5 


28.0 
8 


46.0 
3 


85.0 
4 


250 
63 
1.2 


255 
51 
1.8 


300 
38 


268 
51 
1.5 


900 
60 
0.8 


650 
41 
0.8 


775 
50 
0.8 


2100 
75 

0.8 


1000 
22 
1.2 


3250 
41 
0.4 


1.2 
0.9 
0.1 

2.2 
23 


1.4 
0.3 
0.2 
1.9 
15 


1.6 

0.3 

"L9" 
21 


1.4 
0.5 
0.15 

2.0 
20 


1.4 

0.5 

"i'.Q 

16 


0.7 
0.4 
0.1 
1.1 
29 


1.0 
0.5 
0.1 
1.5 
22 


0.9 
0.6 
0.1 
1.6 
30 


1.7 
1.0 
0.2 
2.9 
17 


1.6 
1.1 
0.1 
2.8 
24 


$126 

8 


$216 
11 


$266 
10 


$203 
10 


$139 

8 




$139 

8 


$190 
14 


$200 
4 




$106 
59 
6.70 


$71 
33 
3.60 


$78 
33 
3.00 


$85 
42 
4.43 


$75 
63 
4.50 


$46 
57 
1.90 


$60 
60 
3.20 


$58 
56 
4.30 


$109 
85 
2.30 


$99 
97 
3.80 


43% 


67% 


59% 


56% 


53%. 


90% 


71% 


42% 


57% 


54% 



53.0 
5 

2120 

46 
0.8 

1.4 

0.9 
0.1 
2.4 

24 

$195 



$89 
79 
2.50 

51% 



FINANCE AND ECONOMICS — COST OF POWER 



853 




Size of Plant, Horse -Power 
Fig. 586. 



475. Elements of Power-plant Design. — The real problem which 
confronts the designing engineer is not so much the selection and 
arrangement of apparatus for a given set of conditions as it is to foresee 
the conditions under which the plant is likely to operate. For this 
reason the plans for the station should be examined and approved by 
an experienced designing engineer, in case expert service is not em- 
ployed at the outset. It is not sufficient to have a mechanically per- 
fect plant, though of course proper installation is of prime importance. 
The choice of fuel, selection of type of prime mover, size of units 
provision for future expansion, and similar factors bear considerable 
weight upon the ecorcmy of operation. Each proposed installation is 



854 



STEAM POWER PLANT ENGINEERING 



likely to be a problem in itself, and though similar plants may be used 
as patterns, each case should be worked out on its own merits. 

The most important factor in the design of a power station is the 
determination of the probable load curve. This refers not only to the 
average yearly load but also to the maximum daily load which is likely 
to occur, the minimum daily load, temporary peak loads, and probable 
future increase. The station load factor and the yearly load factor 
which have such a marked bearing on the cost of operation may be 
closely approximated from the daily load curves. Steam requirements 
for heating and industrial purposes, water supply, and other forms 
of energy requirements should be considered simultaneously with the 
electrical demands since these factors largely influence the choice of 

























































January U, Very darkand cloudy 




A 


bj 


Sv 
























June 20, Brigh 


t 


/ 


7 




\ 


\ 








































1 










v 




















80 




















1 




/ 






\ 


\ 

\ 




































fl 




/ 






\ 


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Fig. 587. Typical Daily Load Curves, Large Apartment Building. 



prime mover. The curves in Figs. 587 to 589 are taken from the daily 
records of large power stations in Chicago and serve to illustrate the 
great variation in the electrical power demands for different days in 
the year. It is quite evident that an equipment based solely upon the 
average yearly requirements may not be adapted to the best economical 
operation. 

The load curves for manufacturing plants may be predetermined 
with a fair degree of accuracy since the power demands for various 
purposes may be readily segregated and analyzed, but with public 
utility concerns and certain classes of isolated stations the problem is 
largely a matter of judgment. Thus, in the case of an industrial 
plant, the power requirements for lighting, manufacturing purposes, 
heating, ventilation, and sanitation may be closely approximated since 
the size of building, exposure, number of floors, and the number of 



FINANCE AND ECONOMICS — COST OF POWER 



855 



elevators afford a definite basis for analysis; but with public utility 
concerns the probable load depends largely upon the business acumen 
of the management in securing customers, the location of the plant 
and future demands. In the latter case the load curve is based chiefly 
upon the experience of similar plants under comparable conditions of 
operation. 



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In any case the greatest care should be exercised in estimating the 
maximum peak load which is likely to occur. High peak loads with 
low daily average necessitate the installation of large machines which 
are idle or operate uneconomically the greater part of the time and 
result in heavy fixed charges. The financial failure of many electric 
light and power plants is directly traceable to the failure to consider 
the influence of maximum peak loads on the ultimate cost of operation. 
In connection with central-station service every customer represents a 
certain investment, regardless of the amount of power used. Even 
should he consume no power, his account would have to be carried on 
the books and a certain amount of equipment would have to be held 
in readiness to serve him. In order that every customer shall incur 
his share of the expense, the expense of the plant must be apportioned 
between the capacity and output costs. The heavier the peak loads 
the greater will be this charge, and, as is the case with many small 
lighting plants where current is used but three or four hours a day, the 
readiness to serve charge becomes excessive and either the station must 
operate at a loss or the unit cost will appear to be prohibitive. 

The curves in Fig. 590 are taken from recording ammeter and re- 
cording steam meter readings of a 200-kilowatt direct-connected and 



856 



STEAM POWER PLANT ENGINEERING 



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Economy. 



FINANCE AND ECONOMICS — COST OF POWER 



857 



a 45-kilowatt belted generator set installed at the power plant of the 
Armour Institute of Technology and serve to illustrate the influence of 
load on economy for very unfavorable conditions. At 8':00 a.m. the 
small unit is started up with initial load of about 150 amperes. As the 
load increases the water rate decreases, as is shown by the curve AB. 





































































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Day of the Month 
Fig. 591. Monthly Load Curves, Combined Heat and Power Plant, Armour Institute 

of Technology. 



At 9:00 a.m. the load is beyond the capacity of the small machine and 
the large unit is put into service. The increased water rate of the large 
unit over the requirements of the smaller is apparent by the sudden 
rise in the water-rate curve. This is due to the fact that the large unit 
is operating at only 20 per cent of its rating, against full load for the 

20,000 



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Jan. Feb. Mar, Apr. May June July Aug. Sept. Oct. Nov. Dec. Jam. 
Month of the Tear 
Fig. 592. Yearly Load Curve showing Influence of Temperature on Coal Consumption, 
Combined Heat and Power Plant, Armour Institute of Technology. 



small one. The fluctuation of the water rate with the load variation 
is clearly shown. Evidently the two units are not of the proper size 
for the particular load conditions illustrated in Fig. 590. During the 
heating months when live steam is necessary for "make-up" purposes 
the unfavorable engine load has little effect on the ultimate economy, 



858 



STEAM POWER PLANT ENGINEERING 



but during the summer months the loss from this cause may be a serious 
one. 

The curves in Figs. 590 to 592 show that during the winter months in 
a combined heat and power plant the fuel requirements may be prac- 
tically uninfluenced by the electrical demands and increase in electrical 
output does not effect an appreciable increase in fuel consumption, but 
the influence of the outside temperature is clearly indicated. 

BIBLIOGRAPHY. COST OF POWER. 

a * i n * t r> d ( 30: 219 Jan. 26, 1909 

Actual Cost of Power, Power < __ _ . n _ _ , ' , nnn 

' (30: 506 March 16, 1909 

Analysis of Industrial Power Costs, Power 33: 950 June 20, 1911 

An Equitable Sliding Scale for Rates for Electrical 

Power, Engineering News 63: 396 April 7, 1910 

A Practical Study of Power Costs, Engineering 

Magazine 39: 230 May, 1910 

Approximate Cost of Gas Power, Power 28: 526 April 7, 1908 

A Schedule of Rates Involving the Investment Cost 

of Reaching the Consumer, Electrical World .... 57: 1562 June 15, 1911 

Central Station vs. Private Plants, Engineering 87 : 288 Feb. 26, 1909 

Comments on Fixed Costs in Industrial Power 

Plants, Proc. Am. Inst, of Eke. Engrs 30: 467 March, 1911 

Comparative Cost of Power Production, Electrical 

Age 40: 63 March, 1909 

Electrical World 53: 792 April 1, 1909 

Electrical Review and Western Electrician 55: 773 ] April 28, 1909 

Comparative Cost of Operating Steam Plants With 
and Without Hydroelectric Auxiliary Power, 
Engineering News 61: 626 June 10, 1909 

Comparative Costs for Three Types of Power 

Plants, Practical Engineer, U. S 14: 645 Oct., 1910 

Cost Data of Power Plant Installation and Opera- 
tion, Engineering Magazine 42: 549 Jan., 1912 

Cost and Depreciation of Steam and Hydroelectric 

Installation, Electrical World 54: 1558 Dec. 30, 1909 

Costs of a Gas Engine and a Combined Steam 

Plant, Engineering Record 60: 273 Sept. 4, 1909 

Cost of Generating and Distributing Electricity for 
Lighting and Power, N. Y. Edison Co., Engineer- 
ing-Contracting 33: 437 May 11, 1910 

Cost of Electricity, Electrical World 56: 265 Aug. 4, 1910 

Cost of Exhaust Steam Heating, Power 36: 788 Nov. 26, 1912 

Cost Figuring in an Electrical Plant, American 
Machinist 36: 223 Feb. 8, 1912 

Cost of Generating Electrical Energy in Steam- 
driven Central Stations of Small and Medium 
Size, Electrical World 57: 549 March 2, 1911 

Cost of Heat, Light and Power, Power 32: 2043 Nov. 15, 1910 

Power 32: 1892 Oct. 25, 1010 



FINANCE AND ECONOMICS — COST OF POWER 



859 



Cost of Industrial Power, Power 

Power 

Proc. Am. Inst, of Elec. Engrs 

Jour. Asso. Eng. Soc 

Cost of Isolated Plant Power, Practical Engineer, 
U.S 

Power 

Iron Age 

Cost of Installing a Modern Steam Turbine Plant, 

Electrical World 

Cost of a Kilowatt-hour, Electrical World 

Cost of Power (General Data), Engineering Record. 

Power 

Practical Engineer, U.S 



Electrical World 



Proc. Am. Inst, of Elec. Engrs 

Cost of Power in Varying Units, Engineering 
Magazine 

Cost of Power in an Office Building Plant, Power . . . 

Cost of Power Producing in a Representative Cen- 
tral Station, Electrical World 

Cost of Power in a 3000-Kilowatt Turbine Plant, 
Electrical Review and Western Electrician 

Cost of Power in a 1500-Kilowatt Central Station, 
Engineering News 

Cost of Power in Four Central Stations, Electrical 
World 

Cost of Power in a 5000-Kilowatt Central Station, 
Power 

Cost of Steam Power in a Mill Plant, Engineering 
Record 

Cost of Steam Power Plants, Electrical Railway 
Journal 

Cost of Steam Power in Canada from Actual Per- 
formances, Canadian Engineer 

Cost Systems and Time Keeping, Engineering News 

Discrimination on Central Station Rates, Engineer- 
ing Magazine 

Economies of Power Supply from Low Head Plants 
and Steam Plants, A Comparison of Costs, Engi- 
neering-Contracting 

Electrical Power Costs in Small Stations, Engi- 
neering Record 

Estimated Cost of an Industrial Plant, Electrical 
World 

Estimating Electric Power Costs, Central Station. . 



32: 501 
33: 839 
30: 485 
44: 150 

16: 183 
35:460 
39:411 

55: 163 
54: 853 
60: 711 
32: 1790 
13:720 
[16: 422 
16: 777 
58: 631 
58: 1283 
59:39 
28:283 

35: 562 
34: 241 

54: 553 

55:62 

61:471 

55: 813 

30: 305 

62: 183 

39: 1081 

21:259 
60: 621 

41:406 



March 15, 1910 
May 30, 1911 
March 10, 1911 
March, 1910 

Feb. 15, 1912 
April 2, 1912 
Feb. 15, 1912 

Jan. 20, 1910 
Oct. 7, 1909 
Dec. 25, 1909 
Oct. 4, 1910 
Dec, 1909 
April 15, 1912 
Aug. 1, 1912 
Sept. 9, 1911 
Nov. 25, 1911 
Jan. 6, 1912 
April, 1909 

July, 1908 
Aug. 15, 1911 

Sept. 2, 1909 

Oct. 2, 1909 

April 29, 1909 

March 31, 1910 

Feb. 9, 1909 

Aug. 13, 1910 

June 22, 1912 

Aug. 31, 1911 
Dec. 3, 1908 

June, 1911 



35: 696 


June 14, 1911 


59:30 


Jan. 9, 1909 


56: 1291 


Dec. 1, 1910 


11:1 


July, 1911 



860 STEAM POWER PLANT ENGINEERING 

First Cost of Plant and Cost of Generating and 

Distributing Electricity for Lights, Brooklyn 

Edison Co., Engineering-Contracting 33: 393 April 6, 1910 

[32:1894 Oct. 25, 1910 

Hotel Power Costs, Power 1 32: 2003 Nov. 8, 1910 

[ 32: 2243 Dec. 20, 1910 

Isolated Station Records and Accounting, Power. . 28: 669 April 28, 1908 

Labor Costs and Load Factors in Central Station 

Practice, Electrical Review and Western Electrician 59:423 Aug. 26, 1911 

Method of Studying Power Costs with Reference to 

the Load Curve and Overload Economies, Proc. 

Am. Inst. Elec. Engineers 



31:1685 July, 1912 

Notes on Factoring Power Costs, Electric Journal. 8: 559 June, 1911 

Operation Cost of Wisconsin's Capitol Plant, Power 36: 592 Oct. 22, 1912 

Operating Costs of Large Units, Power 32: 981 May 31, 1910 

Power Costs for Factories, Engineering Record .... 60: 604 Nov. 27, 1909 

Power Plant Operating Cost, Power 35: 670 May 7, 1912 

Power Plant Waste, Gassier' s Magazine 36: 497 Oct., 1909 

Production Costs in Progressive Central Stations, 

Electrical World 58: 557 Sept. 2, 1911 

Real Theory of Real Electric Rates, General Electric 

Review 14: 157 April, 1911 

Relation of Load Factor to Power Costs, Jour. 

Western Soc. of Engineers 14: 241 April, 1909 

Engineering Record 59: 702 June 5, 1909 

Representative Data from Electric Power Plant 

Operation, Engineering Magazine 36: 833 Feb., 1909 

Some Notes on Isolated Plants, Proc. Am. Inst, of 

. Elec. Engrs Jan. 12, 1912 

Systems of Charging for Electrical Energy (W. T. 

Ryan), Engineering Magazine 37: 47 April, 1909 

The Small Station and its Economical Operation, 

Western Electrician 43: 10 July 4, 1908 

The Valuation of Steam Power Plants (C. T. Main), 

Electrical Age 39: 228 Oct., 1908 

Useful Figures from Practical Power Plant Opera- 
tion, Electrical World 54: 781 Sept. 30, 1909 

Working Results from a Gas-Electric Power Plant 

(J. R. Bibbins), Proc. Am. Inst, of Elec. Engrs 27: 1123 July 1, 1909 



CHAPTER XIX. 



TYPICAL SPECIFICATIONS. 



476. Sample Specifications for a Cross Compound Non-Condensing 
Engine. — For and in consideration of the amount and terms named in the letter 
accompanying this specification, and of the same date, we propose to furnish 
f.o.b. cars at our factory, Elizabethport, N.J., for account of The Armour 
Institute op Technology, Chicago, 111., One Ball & Wood Horizontal 
Four- Valve (Corliss) Center-Crank Engine, designed for direct connection 
to a direct-current generator, as follows : 

General Horse power 350 to 375 

Dimensions. Diameter of cylinders H.P., 17; L.P., 27 inches 

Length of stroke 18 inches 

Revolutions per minute 175 to 200 

Governor wheel diameter, 90; face, 21 inches 

Width of belt (if belted) 20 inches 

Diameter of steam pipe 7 inches 

Diameter of exhaust pipe 10 inches 

Crosshead pins 6 in. long, 8 in. diameter 

Crank pins 9 in. long, 9$ in. diameter 

Main bearings 20 in. long, 9 in. diameter 

Wearing surface of crossheads 242 square inches 

Width of engine over all 14 feet 6 inches 

Length of engine over all 20 feet 3 inches 

Weight complete 60,000 pounds 

Rating. The rated power of the engine specified is based on an initial 

pressure of 120 pounds (measured in the cylinder), cutting off at 
about one third stroke in both cylinders, without vacuum, when 
operating at 200 revolutions per minute. 
Fittings. With each engine is furnished the following complete list of 

fittings: 

One extended shaft (omitted if belted engine), 

One sub-base with extension for dynamo (omitted if belted 

engine), 
One self-oiling outboard bearing (omitted if belted engine), 
One throttle valve, 
One lubricating system consisting of pipes, sight feeds, and 

oil reservoir, 
One cylinder lubricator, nickel plated, 
861 



862 STEAM POWER PLANT ENGINEERING 

Fittings — Continued. 

One set special steel wrenches, 

One socket wrench for piston, 

One socket wrench for connecting rod bolts, 

One steel wrench for hexagon nuts, 

One connecting rod set screw wrench, 

One monkey wrench, 

One spanner for valve stem gland, 

One eye bolt for pillow block cap, 

Two push-off bolts, 

One set grease cups, 

Two oil cups, 

One hand oil pump, 

One set brass oil cups, 

Three nipples for drip pipes, for frame, 

Four nipples for cylinder drips, 3 inches long, 

One nipple for throttle bleeder, 

One globe valve for throttle bleeder, 

Four globe valves for cylinder drips, 

One set foundation bolts, nuts, and plates, 

One template for locating bolts, 

One governor wheel and keys, 

One balance wheel and keys (omitted if direct connected 
engine), 

Packing for piston and valve rods, 

One one-gallon can cylinder oil, 

One one-gallon can engine oil, 

Two cans grease, 

Two wedges for wheels, 
Cylinders. Cylinders are made of hard and close-grained iron, and under 

the influence of oil and wear the walls will rapidly acquire a fine, 
smooth glaze. Radiation is prevented by a thick jacket of 
asbestos cement, outside of which is neatly fitted an orna- 
mented jacket. Openings and globe valves are provided for 
drainage. 
Connecting The connecting rods are of forged machinery steel of low 
Rods. carbon and fitted with heavy straps with keys and bolts for 

adjustment. 
Crank Pin The crank pin boxes are of cast iron lined with Babbitt 

Boxes. metal. 

Crosshead The crosshead boxes are of the best quality of phosphor 

Boxes. bronze. 

Crossheads. The crossheads are of cast iron faced with Babbitt metal on the 
wearing surfaces. The crosshead pins are pressed into the cross- 
heads. 






TYPICAL SPECIFICATIONS 



863 



Piston. The piston rods are of special hammered steel, threaded and 

screwed into the crossheads, and locked fast with special nut 
counterbored at the end to cover threads, finished and case- 
hardened. The other ends of the rods will be fitted to the pistons 
with thread and locked with nut. The pistons will be fitted with 
two rings turned eccentric and cut open at the thinnest part, the 
ends being halved so as to lap when in position. 

Valves. Both the admission and exhaust valves are of the Corliss 

pattern. The former are provided with double ports, and are 
actuated from a wrist plate receiving its motion from the governor 
placed in the fly wheel of the engine. This governor controls the 
valves of both the high and low-pressure cylinders and possesses 
a range of cut-off from to about f stroke. 

The exhaust valves are driven from a wrist plate through an 
adjustable eccentric by which any desired degree of compression 
can be obtained. 

Speed. The use of Corliss valves, arranged as described in the foregoing 

paragraph, permits an increased speed over the common type 
of Corliss engine with releasing gear, and while yielding the 
same economy dispenses with many working parts, and, what 
is more important, with the large and cumbersome fly 
wheel which has so often proved a source of danger in slow- 
speed engines. 

The frame is proportioned for great strength and the metal is 
placed where it is most needed. An oil groove is cast around the 
bottom of the frame to protect the foundation. 

Main bearings are fitted with removable Babbitt shells which 
can be replaced when necessary. Special care is taken to have 
these bearings of ample length to support the wheels and stand 
the strain of power transmission or the weight of armature when 
direct connected. In the latter case a self-oiling outboard bearing 
is provided to carry the outer end of shaft. 

Guides, The guides are known as the locomotive pattern and are inter- 

changeable. They are carefully scraped to surface plates and 
provision made for taking up wear. 

Crank Shaft. The crank shaft is of the best quality of steel, being carefully 
counterbalanced by cast-iron disks in which the necessary weight 
is placed. In direct connected engines this shaft is either 
extended in one piece to carry the armature or made in two 
pieces and coupled. 

Governor. The governor is of the inertia type and has a swinging eccen- 

tric, the eccentric center moving across the end of the shaft about 
an outside point, and giving a lead which varies with the point of 



Frame. 



Main 
Bearings. 



864 



STEAM POWER PLANT ENGINEERING 



Governor 



Balance. 



Oiling. 



Guarantees. 



Drawings. 

Preparation 
for Ship- 
ment. 



Erection. 



Continued. 
cut-off from a maximum, at the latest point, to zero, when the 
governor weights occupy their extreme outward position. Alter- 
ation in speed is obtained by changing the amount of weight in 
the pockets of the lever arm. 

The balance wheel (in the case of belted engine) is made with 
a flanged rim and with a split hub, the hub being secured to the 
shaft by a bolt. The other keys are square, with parallel sides, 
and are inserted without driving. 

The oiling system consists of a simple oil reservoir which sup- 
plies oil through a system of pipes to the points of the engine 
needing lubrication. After fulfilling its functions this oil is drained 
and can be used anew. This does away with the old cumbersome 
oil-cup system and has the great advantage of delivering clean oil 
to the engine. 

Material. We guarantee that the material and workmanship 
are of the best and that all working parts having flat surfaces are 
scraped to surface plates. 

Regulation. That the engine shall regulate within 2 per cent 
under changes of load within the range of the governor, and that 
no reduction of boiler pressure shall reduce the speed until the 
latest point of out-off is reached. 

Steam Consumption. That the steam consumption, when the 
engine is developing its rated power at 125 pounds pressure 
and no vacuum, shall not exceed 22 pounds of dry steam per 
indicated horse power per hour; that the clearance shall not 
exceed 8 per cent. 

With the engine is furnished a drawing showing its details, 
together with foundation plans. 

Every engine is completely erected at our works before ship- 
ment. The castings are rubbed smooth, carefully filled, and the 
engine given two good coats of standard shop color. All bright 
parts are carefully protected against corrosion. The engine is 
dismantled, the small parts being boxed, and in the case of export 
shipment the larger pieces crated. 

Full drawings and directions for erecting the engine will be 
furnished. Template, foundation bolts, nuts, and plates to be 
shipped in advance if necessary, and by freight unless otherwise 
directed. 

If requested we will furnish the services of an expert to superin- 
tend the erection of this engine at the rate of $5 per day added 
to his traveling expenses and board, the purchaser to furnish all 
laboring help. 



TYPICAL SPECIFICATIONS 



865 



Terms. One-half cash on presentation of bill of lading, balance on 

completion of erection. 

The title to the apparatus herein sold shall not pass from The 
Ball and Wood Company until all payments hereunder (including 
deferred payments, if any) shall have been fully made in cash. 
The purchaser agrees to do all acts necessary to perfect and main- 
tain such retention of title in the said Company. All previous 
communications between the parties hereto, verbal or written, are 
hereby abrogated and withdrawn, and this proposal, when duly 
signed and approved, constitutes the agreement between the 
parties hereto, and no modification of this accepted agreement 
shall be binding upon the parties hereto or either of them unless 
such modification shall be in writing, duly accepted by the pur- 
chaser and approved by an executive of the Company. 

Limit. Prices subject to revision after thirty days. Delivery subject 

to strikes, accidents, or causes beyond our control. 



477. Specifications for Horizontal Tubular Steam Boiler.* — The 

following specifications for one 54-inch horizontal return tubular steam 
boiler, pressure 125 pounds, were prepared by the Hartford Steam 
Boiler Inspection and Insurance Company for the Armour Insti- 
tute of Technology, Chicago. 

The boiler to conform to the following conditions and requirements : 

Type and It is to be of the horizontal tubular type, set with overhanging 

General front, and all parts and pieces are to be designed accordingly. 

It is to be 17 feet 2 inches long, outside, and 54 inches in dia- 
meter, measured on the outside of the smallest ring of plates. 
Heads are to be 16 feet inches apart, outside. 

Shell plates are to be three-eighths of an inch thick on the 
edges, of open-hearth fire-box steel, having a tensile strength of 
not less than 55,000 pounds nor more than 62,000 pounds per 
square inch of section, and an elastic limit of not less than half 
the tensile strength, with not less than 56 per cent of ductility, as 
indicated by contraction of area at point of fracture under test, 
and by an elongation of 25 per cent in a length of 8 inches. 

Heads are to be one-half of an inch thick, of best open-hearth 
flange steel. All plates, both of shell and heads, are to be plainly 
stamped with name of maker, brand, and tensile strength ; brands 
so located that they may be seen on each plate after the boiler is 
finished. 

Each shell plate is to bear a coupon which shall be sheared off, 
finished up, and tested by, or for, the maker of the boiler, at his 
expense. Each coupon is to fulfill the foregoing requirements as 



Dimensions. 



Materials : 
Quality, 
Thickness, 
and Tests. 



* Drawings have been omitted. 



866 



STEAM POWER PLANT ENGINEERING 



Materials: Quality, Thickness, and Tests — Continued. 

to strength and ductility, and stand bending down double when 
cold, when red hot, and after being heated and quenched in cold 
water, without signs of fracture. There is not to be more than 
0.035 per cent of sulphur, nor more than 0.035 per cent of phos- 
phorus in the chemical composition of the plates and heads. All 
plates failing to pass these tests will be rejected. All tests and 
inspections of material may be made at the place of manufacture 
prior to shipment. Certified copies of report of tests must be 
sent to the Hartford Steam Boiler Inspection and Insurance 
Company, Hartford, Conn. 

Riveting. The longitudinal seams are to be of the double-riveted butt- 

joint type with double covering strips. They are to be arranged 
to come well above the fire line of the boiler, and break joints 
in the 3 ring courses in the usual manner. The plates are to be 
planed on the caulking edges before rolling. 

All dimensions and proportions are to be shown on accompany- 
ing drawing No. 1502. 

The girth seams are to be of the single-riveted lap-joint type; 
rivets to be of same size as those in longitudinal seams, and 
pitched 2$- inches apart from center to center ; the distance from 
center of rivet to the edge of the plate to be equal to 1J times 
the diameter of rivet hole. 

The rivet holes are to be either drilled in place, or punched at 
least one-quarter of an inch less than full size ; if the latter method 
is used, the plates, after punching, are to be rolled and bolted 
together, and the rivet holes drilled in place one-sixteenth of an 
inch larger than the diameter of the rivets. The plates are then 
to be disconnected. All burrs are to be removed from the edges 
of the holes. Should any holes be in the least 'out of true, they 
are to be brought in line with a reamer or drill; if a drift-pin is 
used for this purpose the boiler will be rejected. 

All rivets are to be driven by hydraulic pressure, wherever 
possible, and allowed to cool and shrink under pressure. This pres- 
sure is to completely fill the rivet holes, producing a tight joint. 

Rivet Ham- The rivets are to be of the best quality of iron or soft steel, 
mer Tests, capable of being hammered flat, when cold, to a thickness of 
one-half their original diameter, or when hot, to one-third their 
original diameter, without showing signs of fracture. In the 
absence of physical test, it is understood that the contractor 
guarantees the above quality of rivets. 
Braces. There are to be 20 braces 1^ inches in diameter in the boiler, 

10 above the tubes on front head, and 10 on rear head, of the 
crow-foot form, arranged as shown on drawing. None of them 
is to be less than 3 feet 6 inches long, and each is to be fastened 



TYPICAL SPECIFICATIONS 



867 



Braces — Continued. 

to shell and heads by two seven-eighths inch rivets at each end ; 
or solid steel, diagonal braces of approved pattern, and of equal 
strength to the former, may be used. Care is to be exercised in set- 
ting them that they may bear uniform tension. Crow-foot braces 
may be flat in body, if of equal strength to those specified above. 
Braces There are to be 4 braces below the tubes in the boiler. Two 

below of these are to be through braces extending from head to head. 

Tubes. Each brace is to be 1 t$ inches in diameter, with a fork formed on 

rear and secured with a 1^-inch turned bolt and nut to a crow- 
foot securely riveted to rear head; these are the inner or central 
braces. The front end of brace is to be upset to a diameter of 1| 
inches, threaded and secured to front head with a nut and washer 
on both the inside and outside of head. 

The 2 remaining braces are each to be 1^ inches in diameter, 
and secured to rear head in same manner as the through braces; 
the front end of the brace is to be extended forward, fitted to side 
of shell, and riveted there with two 1-inch rivets. All to be 
substantially as shown on accompanying diagram of tube head 
No. 2431. 
Tubes: Size, There are to be 36 lap-welded or seamless-drawn tubes, of the 
Number, and best quality with regard to tensile strength and ductility. They 
Arrange- are to be round, straight, and free from all surface defects, prop- 
onent, erly annealed on their ends, and guaranteed by the manufac- 
turers to have been tested to at least five hundred (500) pounds 
per square inch internal hydrostatic pressure. Each tube is to 
be 4 inches in diameter, 16 feet inches long, and not less than 
standard thickness, set in vertical rows, with a clear space 
between them, vertically and horizontally, of 1 inch, except the 
central vertical space, which is to be 2 inches, as shown on accom- 
panying diagram of tube head No. 2431. 

Holes for tubes are to be neatly chamfered off on outside. 
Tubes to be set with a Dudgeon expander, and beaded down at 
each end. Tube holes may be drilled and reamed, or may be 
punched one-quarter inch less than full size, then rose bitted to 
exact diameter. 
Manholes. There are to be two manholes, one 11 x 15 inches, with pressed 

steel frame, double riveted to inside of shell on top, and one 
10 x 15 inches, flanged in front head below tubes, with suitable 
plates, yokes, and bolts, the proportions of the whole such as will 
make them as strong as any portion of the shell of like area. 

Boiler The boiler to be suspended from steel I beams, 6 inches deep, 

Supports. 12^ pounds per foot, by means of eye or U bolts and plate loops. 

There are to be 6 loops, 2 on each side of the boiler, securely 

riveted to boiler shell. The I beams are to be supported on cast- 



868 



STEAM POWER PLANT ENGINEERING 



Boiler Supports — Continued. 

iron columns of square or rectangular section 6 inches square, 
three-quarters inch thick. Each pair of beams is to be connected 
together, 3 inches apart, by tie-bolts and cast-iron separators; one 
separator near each end, and others at intervals of about five 
feet. The top and bottom flanges of columns are to be faced true. 
The whole system of suspension is to be made in the best man- 
ner, properly arranged to allow free expansion of the boiler, 
securely held and supported in every direction, amply strong in 
every part, and finished complete. 

Nozzles. There are to be two heavy cast nozzles, made of gun-iron or 

steel, one 4 inches internal diameter for steam-pipe connection, 
and one 6 inches internal diameter for safety-valve connection, 
each accurately squared on top flange, and securely riveted to 
boiler on top. Forged or pressed steel pipe flanges may be used 
in place of nozzles. 

The flanges of the nozzles to correspond in diameter and thick- 
ness with standard extra heavy pipe fittings. 

Smoke There is to be an opening 10 by 62 inches cut out of front 

Opening. connection on top for attachment of uptake or flue. 

Feed Pipe. There is to be a hole tapped in front head for a brass bushing, 
3 inches above the top of upper row of tubes, and 16 inches from 
center of boiler, on left-hand side, for 1^-inch feed-pipe connection. 
The bushing is to be not less than 2 inches long, to permit both 
the external and internal feed pipes to be screwed into it not 
less than seven-eighths inch. 

Also furnish and put in a 1^-inch feed pipe extending from 
front head back to within two feet of rear head of boiler, thence 
across the boiler to near shell on right-hand side. On this end 
place an elbow with the outlet pointed down as shown on draw- 
ings. Feed pipe is to be properly hung from the braces. 

Blow-off There is to be an extra heavy pressed steel pipe flange, riveted 

Pipe Con- to bottom of shell, near rear end, and tapped to receive a 4-inch 
nection. extra heavy blow-off pipe. Blow-off valve and fittings to be 

extra heavy. 
Fusible There is to be a fusible plug in rear head, two inches above 

Plug. top of upper row of tubes. 

Fittings. There is to be furnished one pop safety valve 3 inches in 

diameter, one 6-inch steam gauge, three three-quarters-inch gauge 
cocks, and one three-quarters-inch gauge glass 12 inches long, all 
to be of approved pattern, and the necessary holes to be made for 
their proper connection. If combination water column is used, 
the steam and water connections between it and the boiler must 
be made by pipe not less that l\ inches in diameter. 



TYPICAL SPECIFICATIONS 



869 



Castings There is to be furnished a substantial cast-iron front, with all 

for Setting, necessary anchor bolts, 10 feet long, closely fitting front connec- 
tion doors with suitable fastening to prevent warping, closely 
fitting furnace doors with liner plates, rear connection door 
16 x 24 inches, with liner plates, grate bars for grate, pattern to 
be selected by purchaser of boiler, 54 inches long by 48 inches wide, 
with suitable bearer bars for same, arch bars for rear connection, 
and all buckstaves, with the necessary bolts or tie rods, and all 
other castings or ironwork of any description necessary for the 
proper construction and setting of the boiler complete. 
In General. The intent of the foregoing specification is to provide for 
material and workmanship of the best quality, and any details of 
equipment not mentioned in this specification, or not shown on the 
drawings, but necessary for the proper completion of the boiler 
ready for operation, and to be hereafter contracted for, must be 
of equally good quality. 

The size and description of parts are to conform substantially 
to the details of the accompanying plan, and the boiler, complete, 
is to be delivered at and all of the material and workman- 
ship is to be subjected to the inspection and approval of the 
Hartford Steam Boiler Inspection and Insurance Company. 

478. Specifications for Barometric Condenser and Auxiliaries. — 

The following sample specifications for a barometric condenser will give 
some idea of the various items called for in the purchase of a condenser 
and appurtenances, the italicized items being specified by the purchaser. 



" We submit herewith our tender for one condensing 

plant as follows: 

Rated The condenser and auxiliary machinery will have sufficient 

Capacity. capacity to condense 250 pounds of steam per minute (equivalent 

to the steam exhausted from engines developing 1000 horse 

power on a basis of 15 pounds of steam per horse power per hour) 

when supplied with cooling water at a temperature of 70 degrees 

Fahrenheit, and maintaining a vacuum at the condenser of 26 

inches of mercury. 

Capacity un- The plant will also have to condense the quantities of steam 

der Variable unc [ er the varying conditions as stated below : 

Conditions. 



Steam Condensed per 
Minute, Pounds. 


Temperature of Cooling 
Water, Degrees F. 


Vacuum Maintained at 

Condenser, Inches of 

Mercury. 


250 
300 
340 
250 
290 


70 
70 
70 
80 
80 


26 
25 
24 
25 
24 



870 



STEAM POWER PLANT ENGINEERING 



Quantity of 

Cooling 

Water. 

Apparatus 
Furnished. 



Price and 
Delivery. 

Terms. 



Superin- 
tendence. 

Steam 
Pressure. 



Head 

Pumped 

against. 

Power 
Consump- 
tion Of 
Auxiliaries. 



The volume of cooling water required when the condenser is 
working under the above conditions will be from 550 gallons to 
650 gallons per minute. 

The apparatus to be furnished by us will consist of: 

One cast-iron condensing vessel, complete with barometric 
tubes and foot valves. 

One automatic vacuum regulator. 

A structural steel framework for supporting the condensing 
vessel. 

One positive rotary pump for supplying the cooling water to 
the condenser. 

One "dry" air pump. 

One horizontal steam engine, arranged to drive the water pump 
by belt and the air pump direct, the latter placed tandem to the 
engine ; the engine is to be fitted with a suitable governor arranged 
for variable speeds. Purchaser to furnish belt. 

Two pulleys, one for the engine and one for the water pump. 

One vacuum gauge. 

Four thermometers. 

We do not include any steam, air, or water pipes, valves nor 
foundation bolts, but will furnish plans showing suitable founda- 
tions and general arrangement of the machinery. 

Our price for one barometric condensing plant, as described, 
including all royalties, and delivered f.o.b. cars our works, is 



Payments as follows: Monthly payments as the work pro- 
gresses in our shops, less 10 per cent. The retained percentage 
to be paid when the condenser is started in service, provided 
this is done within a reasonable time after completion. 

If desired, we will furnish a competent machinist to superintend 
the erection and starting of the plant, charging extra for his serv- 
ices, 50 cents per hour and his traveling and boarding expenses. 

The engine driving the air and water pumps will be capable of 
starting and operating the plant with 1 00 pounds minimum steam 
pressure, and will be built strong enough to 'work under 135 
pounds maximum steam pressure. 

The water pump and engine driving same will be designed to 
raise the injection water from the cold well to the condenser, the 
level of the water in the cold well to be not more than 10 feet below 
the level of the water in the hot well. 

The engine driving the air and water pumps will require approxi- 
mately 2 per cent of the main engine steam when operating under 
rated load and with conditions as above stated. 

The rotary pump has a positive displacement (Bibus' patent) 



TYPICAL SPECIFICATIONS 871 

Power Consumption of Auxiliaries — Continued. 

and is of substantial construction. All the power for propelling 
the water is transmitted through the main shaft, the office of the 
gears being simply to keep the sealing runner in time with the 
propelling runner. Stuffing boxes are placed between the pump 
chamber and bearings to prevent any grit in the water coming 
into contact with the journals. 

The engine and air pump is of the crank and fly wheel type, 
the engine being fitted with a suitable governor arranged for vari- 
able speeds, and the air pump with the patented slide valve. 

479. Specification for Steam, Exhaust, Water, and Condenser Piping 
for an Electric Power Station.* — It is intended that this specification 
shall cover the complete installation of steam piping, exhaust piping, 
injection and discharge piping, drain, drip, blow-off, and boiler-feed 
piping, water piping, valves, separators, anchors, fittings, etc., sub- 
stantially as shown on the accompanying drawing or hereinafter 
described. 

All the materials used throughout must be the best of their respec- 
tive kinds, subject to the inspection and approval of the engineer of 
the purchaser. 

The entire work provided for in this specification is to be constructed 
and finished in every part in a good, substantial, and workmanlike 
manner, according to the accompanying drawings and this specifica- 
tion to the full intent and meaning of the same, and everything necessary 
for the proper and complete execution of the plans and drawings, 
whether the same may have been herein particularly specified or not, 
or indicated in the plans referred to, to be done and furnished in a 
manner corresponding with the rest of the work as well, as truly, and 
as faithfully as if the same were herein particularly described and 
specifically provided for. 

The engineer shall have full power at any time during the progress 
of the work to reject any materials he may deem unsuitable for the pur- 
pose for which they are intended, or which are not in strict conformity 
with the spirit of this specification. He shall also have the power to 
cause any inferior or unsafe work to be taken down and altered at the 
cost of the contractor. 

It is to be understood that the final inspection and acceptance of 
the work are to take place at the building after erection, and that any 
inspection and acceptance of material and workmanship at the mills 9 
shops, etc., to facilitate the progress of the work, shall not preclude 
rejection at the building if the same be unsuitable. 

* Stevens Indicator, Vol. 18, p. 373, 1901. 



872 STEAM POWER PLANT ENGINEERING 

Any disagreement or difference between the purchaser and the con- 
tractor, upon any matter or thing arising from this specification or the 
drawings which form a part thereof, shall be referred to the engineer, 
whose decision and interpretation of the same shall be considered final, 
conclusive, and binding upon both parties. 

Risk and Blame. — The contractor is to assume all risks and bear 
any loss occasioned by neglect or accident during the progress of the 
work until the same shall have been completed and accepted by the 
engineer. 

Permits. — The contractor must pay for all permits and inspec- 
tors' fees or any other charges from borough, city, county, or state 
officers. 

Dimensions. — During the progress of the work the contractor will 
be required to keep at the building site a complete set of drawings 
and a copy of this specification. The contractor will consider dimensions 
shown on drawings to be approximate as to location of machinery,- 
but sufficiently accurate for the purpose. Absolute dimensions must 
be gotten from the location of the boilers, engines, and other machinery 
after they are set. No drawing shall be scaled. 

The purchaser reserves the right to put other parties at work on the 
premises erecting machinery and doing other work during the con- 
tinuance of this contract. 

The contractor must conform to such rules and regulations as are 
in force on the property and carry out this contract with as little 
interference as possible with the other work of the purchaser. 

Drawings.* 

The following is a list of the drawings which accompany this 
specification and which form a part thereof: 

(Title) No. (Drawing Number) 

do do etc. 

Description of Plant. 

Steam Piping. — • The plant consists of four (4) 250-horse-power 
boilers from which steam will be led to two (2) independent 16-inch 
steam headers. Two (2) 8-inch take-offs will be carried from each 
boiler, one (1) to each header. These 8-inch take-offs will consist of 
large radius wrought-iron bends of a quality hereinafter described. 
From each header a 6-inch supply will be led to the low-pressure 
cylinders of these engines. From the header nearest the engine room 
two (2) 6-inch supplies will be led through the partition wall into the 

* Drawings have been omitted. 



TYPICAL SPECIFICATIONS 873 

cellar for the purpose of furnishing steam to the condenser pumps. 
From one (1) header a 6-inch and 5-inch supply will be carried to a 
tandem compound engine and thence 5-inch to the two (2) exciter 
engines. 

The condenser steam lines will be continued under the engine-room 
floor so as to form a reserve steam supply to the exciter engines. 
From the end of each 16-inch main steam header a 3-inch loop will be 
carried to and from the fire and boiler feed pumps. The general 
arrangement with sizes of pipes, separators, valves, etc., is shown on 
drawing No. . 

The contractor's high-pressure steam work will begin with the steam 
nozzles on the boilers and end with the throttle valves of pumps and 
engines. The contractor will furnish all pump throttle valves but not 
engine throttle valves. 

Condenser and Exhaust Piping. — The three (3) jet condensers and 
air pumps will be set under engine-room floor. Injection water will 
be led to the condenser through a 16-inch cast-iron pipe which begins 
at a point 6 feet outside the wall of the boiler house and runs from 
this point to the condensers. After leaving the connections for the 
fourth condenser, this main reduces to 14 inches in diameter; after 
leaving the third it reduces to 10 inches in diameter; and after 
leaving the second it reduces to 3 inches in diameter. 

This injection water main will be connected with the city supply by 
a 5-inch cast-iron water pipe; and the discharge pipe will be of cast- 
iron and will be led from the condensers, beginning with a diameter of 
5 inches, increasing to 10 inches, and again increasing to 14 inches, 
and then to 16 inches after the last condenser, from whence it will be 
carried to a point 6 feet outside of the boiler-house wall. 

The low-pressure cylinder of the tandem compound engine will be 
connected by a 10-inch and a 12-inch pipe with No. 1 condenser. 
Connection will be made between both high- and low-pressure cylin- 
ders of the 750-horse-power engines with 18-inch condenser exhaust 
headers. Between the exhaust header and the cylinder of the engines, 
connections will be made with a free exhaust main leading to and 
from exhaust riser extending through the roof of the boiler house. 
The general arrangement and the sizes of the gate valves, free exhaust 
valves, etc., are shown on drawing No. . 

The exhaust from the two (2) exciter engines and the three (3) 
condensers, the boiler feed pumps, and the fire pumps will be collected 
and led to a 1500-horse-power open-type feed- water heater located in 
the boiler room. This heater will be provided with an exhaust riser 
and exhaust head extending through the roof of the boiler house. 



874 STEAM POWER PLANT ENGINEERING 

The arrangement of the auxiliary exhaust pipe is shown on drawing 
No. . 

Boiler Feed Water Piping. — To the feed-water heater above 
mentioned a 2^-inch supply of fresh water will be furnished. The 
boiler feed pumps will draw the feed water from the heater through a 
6-inch cast-iron suction pipe. They will deliver it through 3-inch 
brass pipes to two (2) hot-water meters equipped with by-passes and 
thence through a 3-inch brass pipe line to boilers. A reserve wrought- 
iron 3-inch feed pipe line will be run parallel to the brass feed pipe 
line and so connected up to it that it may be used as a reserve in case 
of accident to the brass line. Connections from 3-inch lines to boilers 
will be 2J-inch brass pipe. 

The general arrangement of the feed pipe, including gate valves, 
check valves, sizes, etc., is shown on drawing No. ■. 

Drip Piping. — There will be two (2) systems of drip piping, 
including traps, trap by-passes, valves, etc. 

The high-pressure system comprises all the drips from the main 
steam header, separators, and high-pressure steam lines. The water 
from these traps will be led to the feed-water heater, drawing 
No. . 

The low-pressure system comprises all the drips from exhaust mains, 
receivers, and cylinder drain cocks. 

This water will be led to a catch-basin from whence it will be 
pumped by a tank pump. This pump will be furnished by the pur- 
chaser, drawing No. . 

Blow-off Pipes. — A separate and distinct wrought-iron, 2|-inch 
blow-off pipe will be run from the blow-off valve of each of the 
four (4) water-tube boilers. These pipes will be run in trenches in 
the boiler-room floor to catch-basins located outside of building. The 
purchaser will construct the trenches and catch-basins. 

Water Piping. — The purchaser will make a 6-inch connection from 
the city main to the northwest corner of the boiler room. From this 
point the contractor will run 6-inch cast-iron pipe to the inlet of 
the fire pump. From this he will run 6-inch pipe to the several fire 
hydrants. He will connect this 6-inch line with the injection water 
main with a 5-inch cast-iron connection. Near the fire pump he will 
run a 2J-inch connection to the feed-water heater. The pipe around 
the fire pump will be so arranged that ordinarily city pressure will be 
maintained on the water system, but in case of need the fire pump may 
be used to increase this pressure. 

This piping is all shown on drawing No. . 

Wr ought-Iron Pipe. — All wrought-iron pipes or steam lines to be 



TYPICAL SPECIFICATIONS 



875 



guaranteed full-weight lap-welded wrought-iron pipe, in accordance 
with the standard dimensions as given in the table below. But the 
16-inch headers will be 16 inches O.D., 15.04 inches I.D., and shall 
weigh 87 pounds per foot. 



Nominal Internal. 


Actual External. 


Actual Internal. 


Nominal Weight per 
Foot. 


Inches. 


Inches. 


Inches. 


Pounds. 


i 


1.315 


1.048 


1.663 


H 


1.66 


1.38 


2.444 


H 


1.9 


1.611 


2.678 


2 


2.375 


2.067 


3.609 


2* 


2.875 


2.468 


5.739 


3 


3.5 


3.067 


7.536 


3* 


4 


3.548 


9.001 


4 


4.5 


4.026 


10.665 


4* 


5 


4.508 


12.34 


5 


5.563 


5.045 


14.502 


6 


6.624 


6.065 


18.762 


8 


8.625 


7.982 


28.35 


10 


10.75 


10.019 


40.065 


14 


15 


14.25 


57.893 


15 


16 


15.25 


71.77 



All pipes carrying hot water, such as feed-water, blow-offs, and 
high-pressure drip pipes, are to be extra heavy. 

AH lap-welded pipe shall be proved to 500 pounds pressure per 
square inch before shipment. Butt- welded pipe shall be proved to 
300 pounds pressure per square inch before shipment. 

Exhaust pipes are to be wrought iron or steel in every case; sizes 
8 inches and above may be light O.D. tubing with flanges peened 
or expanded on, but must be absolutely tight at 28 inches vacuum. 
These pipes must be tested to 25 pounds hydrostatic pressure after 
erection. 

All high-pressure wrought-iron pipe when in place is to be tested at 
250 pounds air pressure, and must be good for a working pressure of 
175 pounds. All high-pressure steam pipe to be tested after erection 
at 175 pounds steam pressure, to the satisfaction of the engineer in 
every particular. 

Cast-iron Pipe. — The cast-iron pipe for injection and discharge 
water and for all other purposes, as shown on the accompanying 
drawings, shall be made of tough gray iron not less than f inch thick 
at any point, free from blowholes, true to pattern, and of workman- 
like finish. It shall be tested at 250 pounds pressure before erection, 
and when in place shall be tested to 25 pounds hydrostatic pressure. 

Cast-iron pipes inside of buildings are to be flanged with flanged 
fittings. 



876 



STEAM POWER PLANT ENGINEERING 



Outside of buildings they shall be bell and spigot ends, and must be 
coated with tar both inside and outside. 

Valves. — All steam valves over two inches, except throttle valves, 
will be gate valves. All gate valves will be Chapman's best make or 
equivalent. 

On high-pressure steam lines, valves must be fitted with bronze re- 
movable seats, outside screw and yoke, and by-passed from 5 inches up. 

On exhaust lines, standard pattern gate valves will be used (Chap- 
man or equivalent), with inside screw. 

Globe valves on high-pressure steam and water lines will be Schutte's 
make or equivalent. 

Free exhaust valves will be either Schutte's make or equivalent. 

Exhaust valves between engines and condensers, injection water 
valves at condensers, as well as priming water valves, will have their 
stems extended and fitted with floor stands, so that they can be 
operated from floor above. Floor stands will be polished all over, 
with polished hand wheels and indicators. 

All high-pressure valves will be tested to the satisfaction of the 
engineer, at a hydrostatic pressure of 350 pounds per square inch 
and air pressure of 100 pounds between gates. 

Standard valves for exhaust and water will be tested to 150 pounds. 

Flanges and Fittings. — All flanges and fittings for high-pressure 
steam lines 3 inches and over will be open-hearth steel castings. 
These castings must be perfectly solid, made from heavy patterns, and 
free from blowholes or other defects. They must be tested at works 
to 500 pounds hydrostatic pressure, and guaranteed for a working 
steam pressure of 175 pounds per square inch. 

The diameter, thickness, and drilling of flanges must not be less than 
the dimensions given in the following table. 



Size of Pipe. 


Diameter of 
Flange. 


Thickness. 


Number of 
Bolts. 


Size of Bolts. 


Inches. 


Inches. 


Inches. 




Inch. 


16 


25 


1! 


20 


1 


14 


23 


if 


16 


1 


12 


20 


l* 


16 


1 


10 


174 


if 


12 


1 


9 


16 


l* 


12 


1 


8 


15 


H 


12 


7 
8 


7 


14 


H 


12 


I 


6 


13 


1* 


8 


f 


5 


11 


H 


8 


I 


4* 


104 


1* 


8 


1 


4 


10 


i* 


8 


3 

i 


H 


9 


l 


8 


f 


3 


9 


l 


8 


f 



TYPICAL SPECIFICATIONS 877 

Below 3 inches, on all high-pressure steam, water, or drip lines, extra 
heavy screwed cast-iron fittings are to be used, with sufficient extra 
heavy flange unions, so that any section of pipe can be readily taken 
out without disturbing the balance. 

The brass boiler feed piping will be made up with extra heavy brass 
screwed fittings made from cast-iron patterns. 

Pipe will be iron pipe size. 

Brass flange unions will be made from standard cast-iron patterns, 
and a sufficient number used to make up the sections readily. 

On low-pressure piping, standard fittings and flanges are to be used. 
Sizes 6 inches and above are to be flanged, and below 6 inches, screwed, 
with sufficient number of flange unions in same. 

All high-pressure flanges are to be recessed on face, pipe is to be 
screwed, then peened up tight, and must stand the tests given 
above. 

Standard flanges must stand the tests as stated. 

All bolt holes must be drilled in solid metal; no cored holes will be 
allowed. 

On blow-off lines, long sweep fittings are to be used, with extra 
heavy flange unions. 

Gaskets and Bolts. — All gaskets on high-pressure steam lines must 
be copper. 

On exhaust and water lines metallic gaskets are to be used. 

All bolts must have hexagonal heads and nuts, no matter what size, 
points to be finished and nuts to be cold punched. 

Pipe Covering. — All high-pressure steam lines, separators, and all 
drip lines between high-pressure steam lines and their traps will be 
covered with an approved magnesia pipe covering. All fittings will 
have molded magnesia coverings to fit them neatly. 

Exhaust Heads. — The contractor will furnish and erect two (2) 
exhaust heads, one on the free exhaust from the engines and the other 
on the free exhaust from the feed-water heater. 

Pipe Supports. — There will be furnished -and erected the necessary 
wrought-iron hangers to support the two (2) 16-inch headers in the 
boiler room. There will also be supplied the necessary supports for 
the steam mains to the pumps and exciter engines. The injection 
water main and the condenser exhaust header will be hung from the 
engine room floor beams. The free exhaust header and the discharge 
water main will be supported on piers and saddles from the cellar 
floor. The saddles to be furnished by the contractor. 

Separators. — The contractor will furnish and erect, as shown on 
drawing No. , eleven (11) steam separators. 



878 



STEAM POWER PLANT ENGINEERING 



Anchors. — All live-steam mains, and especially the 16-inch steam 
headers in the boiler room, will be anchored so as to secure no move- 
ment (at point of anchorage) on account of the expansion and con- 
traction or vibration. Anchors to be placed where needed after the 
rest of the work is completed. 

Relief Valves. — The boiler-feed and fire pumps shall be fitted 
with automatic controlling devices so that a certain definite maxi- 
mum pressure may be maintained on the feed-water and fire 
systems. 

Overflow Pipes. — The contractor will furnish an overflow relief valve 
for the fire pump and will connect same with the drain outside the boiler 
house. He will also connect the blow-off and overflow from the feed- 
water heater with the said drain. 

Long Radius Bends. — Where shown on the drawings, changes in 
direction of pipe runs will be made with long radius bends. These 
bends will be of wrought-iron of the quality described above. Nine 
inches of straight pipe will be left on the ends of the bends to cut 
necessary threads. 

The following table shows the minimum radii for these wrought-iron 
bends. 



Size. 


Minimum Radius. 


Size. 


Maximum Radius. 


Inches. 


Inches. 


Inches. 


Feet. Inches. 


2 


4* 


7 


3 


2h 


6 


8 


3 4 


3 


8 


9 


4 


3i 


10 


10 


4 4 


4 


14 


12 


5 6 


4i 


16 


14 


7 


5 


20 


16 


7 6 


6 


24 













I- Beam Supports. — The contractor will furnish and erect on the 
top chord of the roof trusses in the boiler room a sufficient length of 
I beams or channels from which he will hang the 16-inch headers. 
The steam connections between 16-inch header and the engines will be 
supported by underneath rod trusses, and vibration will be taken up 
by means of lateral ties. 

Flange Covering. — After all the other work is completed and the 
plant has been in operation not less than two (2) weeks, the con- 
tractor will cover all flanges and other joints with an approved 
magnesia covering. 



TYPICAL SPECIFICATIONS 879 

Reducing and Relief Valves. — The contractor will furnish and erect 
two (2) reducing valves on the throttle valves of the low-pressure 
cylinders of the 750-horse-power engines. These valves will be of 
the Kiely type or equivalent. 

There will also be furnished and erected the necessary traps from 
the receivers between the high and low-pressure cylinders. 

Check Valves. — All drip lines will be equipped with check valves 
which will be set at the lowest point in the line. Two (2) three-inch 
brass checks will be placed on the boiler feed pump outlets, all as 
shown on drawing No. . 

The outlet of the boiler feed pumps will also have a relief valve 
which will return the water to the feed-water heater in case the pres- 
sure on the pump outlet should rise above a certain definite point. 

Traps. — All traps on high-pressure drip lines will be extra heavy 
steam traps. 

Painting. — All pipes shall be painted with two (2) coats of slate 
graphite and linseed oil, or other good pipe paint satisfactory to the 
engineer. 

The exhaust and condenser piping must be thoroughly painted twice 
under vacuum. 

Damper Regulator. — The purchaser will furnish the damper regu- 
lator, but the contractor will set it and connect it with both steam 
headers, water supply, and damper. 

Shields. — Where steam mains pass through the partition wall of 
the building the contractor will neatly close the opening with a sheet- 
metal shield so as to prevent dust from the boiler room from entering 
the engine room. 

Meters. — The contractor will furnish and connect as shown on 

drawing No. two (2) 3-inch hot-water meters, approved by the 

engineer, to have a capacity of 200 gallons per minute. 

Gauges. — The purchaser will furnish the customary gauges and 
boards, which will be set up and connected by the contractor. 

Machinery. — The purchaser will furnish all engines, pumps, con- 
densers, feed-water heater, boilers, foundations, but the contractor 
shall connect up the above machinery according to the evident intent 
and meaning of this specification. 

The purchaser will not furnish any pipe, valves, fittings, pipe cover- 
ing, etc., or anything connected with the work except the machinery 
mentioned above. 



880 STEAM POWER PLANT ENGINEERING 

480. Government Specification and Proposal for Supplying Coal. 

U. S. Treasury Department. 

United States 

, 190.. 

PROPOSAL. 

1 Sealed proposals will be received at this office until 2 o'clock p. m., 

2 , 190 . . , for supplying coal to the United States 

3 building at 

4 as follows : 

5 

6 

7 

8 The quantity of coal stated above is based upon the previous annual 

9 consumption, and proposals must be made upon the basis of a delivery of 

10 10 per cent more or less than this amount, subject to the actual requirements 

11 of the service 

12 Proposals must be made on this form, and include all expenses incident 

13 to the delivery and stowage of the coal, which must be delivered in such 

14 quantities, and at such times within the fiscal year ending June 30, 190 , 

15 as may be required. 

16 Proposals must be accompanied by a deposit (certified check, when 

17 practicable, in favor of ) 

18 amounting to 10 per cent of the aggregate amount of the bid submitted, as 

19 a guaranty that it is bona fide. Deposits will be returned to unsuccessful 

20 bidders immediately after award has been made, but the deposit of the 

21 successful bidder will be retained until after the coal shall have been 

22 delivered, and final settlement made therefor, as security for the faithful 

23 performance of the terms of the contract, with the understanding that the 

24 whole or a part thereof may be used to liquidate the value of any deficiencies 

25 in quality or delivery that may arise under the terms of the contract. 

26 When the amount of the contract exceeds $10,000, a bond may be exe- 

27 cuted in the sum of 25 per cent of the contract amount, and in this case, the 

28 deposit or certified check submitted with the proposal will be returned after 

29 approval of the bond. 

30 The bids will be opened in the presence of the bidders, their representa- 

31 tives, or such of them as may attend, at the time and place above specified. 

32 In determining the award of the contract, consideration will be given to 

33 the quality of the coal offered by the bidder, as well as the price per ton, 

34 and should it appear to be to the best interests of the Government to 

35 award the contract for supplying coal at a price higher than that named in 

36 lower bid or bids received, the award will be so made. 

37 The right to reject any or all bids and to waive defects is expressly 

38 reserved by the Government. 



TYPICAL SPECIFICATIONS 881 



DESCRIPTION OF COAL DESIRED.* 

39 Bids are desired on coal described as follows : 

40 , 

41 

42 

43 

44 

45 

46 

47 

48 

49 

50 Coals containing more than the following percentages, based upon dry 

51 coal, will not be considered: 

52 Ash per cent. 

53 Volatile matter per cent. 

54 Sulphur per cent. 

55 f Dust and fine coal as delivered at point of consumption per cent. 



DELIVERY. 

56 The coal shall be delivered in such quantities and at such times as the 

57 Government may direct. 

58 In this connection, it may be stated that all the available storage capacity 

59 of the coal bunkers will be placed at the disposal of the contractor to 

60 facilitate delivery of coal under favorable conditions. 

61 After verbal or written notice has been given to deliver coal under this 

62 contract, a further notice may be served in writing upon the contractor to 

63 make delivery of the coal so ordered within twenty-four hours after receipt 

64 of said second notice. 

65 Should the contractor, for any reason, fail to comply with the second 

66 request the Government will be at liberty to buy coal in the open market, 

67 and to charge against the contractor any excess in price of coal so purchased 

68 over the contract price. 

SAMPLING. 

69 Samples of the coal delivered will be taken by a representative of the 

70 Government. 

71 In all cases where it is practicable, the coal will be sampled at the time 

* Note. — This information will be given by the Government as may be deter- 
mined by boiler and furnace equipment, operating conditions, and the local market. 
f Note. — All coal which will pass through a |-inch round-hole screen. 



882 STEAM POWER PLANT ENGINEERING 

72 it is being delivered to the building. In case of small deliveries, it may be 

73 necessary to take these samples from the yards or bins. The sample 

74 taken will in no case be less than the total of one hundred (100) pounds, to 

75 be selected proportionally from the lumps and fine coal, in order that it 

76 will in every respect truly represent the quality of coal under considera- 

77 tion. 

78 In order to minimize the loss in the original moisture content the gross 

79 sample will be pulverized as rapidly as possible until none of the fragments 

80 exceed | inch in diameter. The fine coal will then be mixed thoroughly 

81 and divided into four equal parts. Opposite quarters will be thrown out, 

82 and the remaining portions thoroughly mixed and again quartered, throw- 

83 ing out opposite quarters as before. This process will be continued as 

84 rapidly as possible until the final sample is reduced to such amount that all 

85 of the final sample thus obtained will be contained in the shipping can or 

86 jar and sealed air-tight. 

87 The sample will then be forwarded to the Chief Clerk of the Treasury 

88 Department, care of the storekeeper. 

89 If desired by the coal contractor, permission will be given to him, or his 

90 representative, to be present and witness the quartering and preparation of 

91 the final sample to be forwarded to the Government laboratories. 

92 Immediately on receipt of the sample, it will be analyzed and tested by 

93 the Government, following the method adopted by the American Chemical 

94 Society, and using a bomb calorimeter. A copy of the result will be mailed 

95 to the contractor upon the completion thereof. 

CAUSES FOR REJECTION. 

96 A contract entered into under the terms of this specification shall not 

97 be binding if, as the result of a practical service test of reasonable duration, 

98 the coal fails to give satisfactory results due to excessive clinkering, or to 

99 a prohibitive amount of smoke. 

100 It is understood that the coal delivered during the year will be of the 

101 same character as that specified by the contractor. It should, therefore, 

102 be supplied, as nearly as possible, from the same mine or group of mines. 

103 Coal containing percentages of volatile matter, sulphur, and dust higher 

104 than the limits indicated on line 54, and coal containing a percentage of 

105 ash in excess of the maximum limits indicated in the following table will 

106 be subject to rejection. 

107 In the case of coal which has been delivered and used for trial, or which 

108 has been consumed or remains on the premises at the time of the deter- 

109 mination of its quality, payment will be made therefor at a reduced price 

110 computed under the terms of this specification. 

111 Occasional deliveries containing ash up to the percentage indicated in 

112 the column of "Maximum limits for ash," on page 700, may be accepted. 



TYPICAL SPECIFICATIONS 



883 



113 Frequent or continued failure to maintain the standard established by 

114 the contractor, however, will be considered sufficient cause for cancellation 

115 of the contract. 

* PRICE AND PAYMENT. 

116 Payment will be made on the basis of the price named in the proposal 

117 for the coal specified therein, corrected for variations in heating value and 

118 ash, as shown by analysis, above and below the standard established by 

119 contractor in this proposal. For example, if the coal contains two (2) 

120 per cent, more or less, British thermal units than the established standard, 

121 the price will be increased or decreased two (2) per cent accordingly. 

122 The price will also be further corrected for the percentages of ash. For 

123 all coal which by analysis contains less ash than that established in this 

124 proposal a premium of 1 cent per ton for each whole per cent less ash will 

125 be paid. An increase in the ash content of two (2) per cent over the 

126 standard established by contractor will be tolerated without exacting a 

127 penalty for the excess of ash. When such excess exceeds two (2) per cent 

128 above the standard established, deductions will be made from price paid 

129 per ton in accordance with following table : 



Ash as estab- 
lished in 
proposal. 


No 

deduc- 
tion for 
limits 
below. 


Cents per ton to be deducted. 


Maxi- 


2 


4 


7 


12 


18 


25 


35 


mum 
limits 




Per 


centages 


of ash in dry coal. 






ash. 


Per cent. 
5 


7 
8 
9 

10 
11 
12 
13 
14 
15 
16 
17 
18 
19 
20 


7- 8 

8- 9 
9-10 

10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 


8- 9 
9-10 
10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 
21-22 


9-10 
10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 


10-11 
11-12 
12-13 
13-14 
14-15 
15-16 


11-12 
12-13 
13-14 


12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 
21-22 


13-14 
14-15 
15-16 
16-17 
17-18 


12 


6 


13 


7 


14 


8 


14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 
21-22 
22-23 


14 


9 


15 


10 


16 


11 


16-17 
17-18 
18-19 
19-20 
20-21 
21-22 
22-23 


16 


12 


17 


13 


18 


14 


19 


15 


19-20 
20-21 
21-22 
22-23 


19 


16 






20 


17 






21 


18 








22 















* Note. — The economic value of a fuel is affected by the actual amount of com- 
bustible matter it contains, as determined by its heating value shown in British 
thermal units per pound of fuel, and also by other factors, among which is its ash 
content. The ash content not only lowers the heating value and decreases the 
capacity of the furnace, but also materially increases the cost of handling the coal, 
the labor of firing, and the cost of the removal of ashes, etc. 



884 STEAM POWER PLANT ENGINEERING 

Proposals to receive consideration must be submitted upon this form and contain 
all of the information requested. 



, 190 

The undersigned hereby agree to furnish to the U. S 

building at , the coal described, in tons 

of 2,240 pounds each and in quantity, 10 per cent more or less than that stated 
on page 697, as may be required during the fiscal year ending June 30, 190 , 
in strict accordance with this specification; the coal to be delivered in such 
quantities and at such times as the Government may direct. 

Price per ton (2,240 pounds) $ 

Commercial name of the coal 

Name of the mine or mines 

Location of the mine or mines 

Name or other designation of the coal bed or vein 

Size (indicate information which will apply) — 

Unsized Lump Run of mine 

C Round ) ~ 

Screened, through inch and over. . . . .inch -j Square ) 

I Bar screen. 
Data to establish a basis for payment: 

British thermal units in coal as delivered 

Ash in dry coal (Method of American Chemical Society) per cent. 

It is important that the above information does not establish a higher standard than can be 
actually maintained under the terms of the contact; and in this connection it should be noted 
that the small samples taken from the mine are invariably of higher quality than the coal actually 
delivered therefrom. It is evident, therefore, that it will be to the best interests of the contractor 
to furnish a correct description with average values of the coal offered, as a failure to maintain the 
standard established by contractor will result in deductions from the contract price, and may 
cause a cancellation of the contract, while deliveries of a coal of higher grade than quoted will be 
paid for at an increased price. 



Signature : 

Address : 



Name of corporation, , 

Name of president, 

Name of secretary, 

Under what law (State) corporation is organized:. 






CHAPTER XX. 

TYPICAL CENTRAL STATIONS. 
Steam Turbines. — Alternating Currents. 

Fisk Street Station. — The Fisk Street Station of the Commonwealth 
Edison Company, Chicago, is an excellent example of modern central- 
station practice. The present (November, 1912) rated capacity of 
the plant is 120,000 kilowatts, though space is available for a consider- 
able increase. The station is located on the banks of the Chicago River 
near Fisk and Twenty-second streets, as indicated in Fig. 593, and is 
about one and one-half miles south of the center of distribution of the 
present load. The location of the station between the east and west 
slips of the river secures an unusual advantage in the location of the 
intake and discharge tunnels, and the extent of property affords ample 
storage capacity for coal. The Chicago, Burlington & Quincy railway 
extends into the property, giving excellent facilities for the transporta- 
tion of coal, ashes, construction materials, and machinery. The plant 
is constructed on the unit basis, each turbine and generator having 
its own boilers, auxiliaries, and piping system, thus permitting any unit 
to be shut down without interfering with the operation of the rest of 
the system. 

Building. — The main building rests on piles, driven to hard pan, 
capped with a grillage of I-beams and concrete. The walls are of red 
pressed brick trimmed with white Bedford stone. The windows are 
25 feet wide and 32 feet high, the sections of which are operated by 
compressed air. Fig. 593a gives a view of the north elevation. Large 
skylights afford ample light and ventilation. The entire interior wall 
surface of the turbine room is finished with white enameled brick 
trimmed with terra cotta. The boiler-room walls have an eight-foot 
wainscoting of enameled brick, the remainder being red pressed brick. 
The floors are of concrete, that in the turbine room being covered with 
two-inch hexagonal terra-cotta tile. The roofs are of Roebling concrete. 
The total width of the building is 243 feet, the turbine room taking 
up 61 feet, the boiler room 142 feet, and a car track the remainder. 
Two motor-driven cranes span the turbine room and run the full length 
of the building. A five-ton auxiliary hoist is also provided on the main 

885 



886 



STEAM POWER PLANT ENGINEERING 



5 
•a 

I 
g 







TYPICAL CENTRAL STATIONS 



887 



cranes. In the boiler room a small hand-power crane serves each two 
batteries of boilers. 

Coal and Ash Handling. — An 
interior shed extends the entire 
length of the east end of the build- 
ing, as indicated in Figs. 594 and 
595. Coal is brought in on cars 
and dumped or shoveled into a 
track hopper, from which it is de- 
livered to the overhead bunkers by 
the conveying system. A crusher 
is placed between the track hopper 
and main conveyor to be used in 
case lump coal is furnished. These 
bunkers have a capacity of 1200 
tons each, sufficient for several 
days run. The conveyors are 
driven by a 15-horse-power motor 
and are of the McCaslin pattern, 
endless chain, with overlapping 
buckets, each bucket having a 
capacity of 100 pounds. The con- 
veyors move at a variable speed 
giving a service capacity up to 75 
tons per hour for each unit. A 
separate conveyor and bunker is 
installed in each section of 8 boilers. 
The coal bunkers feed through 
flexible down spouts to the stoker 
magazines. Underneath the front 
end of the stoker is a fine-coal 
hopper which collects the fine coal 
falling through the grate and 
discharges it into the conveyor 
system, as in Fig. 97. The ashes 
collect in the ash pit, from which 
they are dumped into the conveyor 
and carried to an ash bin directly 
over the coal track. Illinois screen- 
ings furnish the greater part of 
the fuel. Provision is also made for outside storage. 

Boilers. — The boiler plant is divided into five sections, each section 




888 



STEAM POWER PLANT ENGINEERING 




TYPICAL CENTRAL STATIONS 



889 




890 STEAM POWER PLANT ENGINEERING 

consisting of sixteen 500-horse-power B. & W. boilers arranged in bat- 
teries of eight and equipped with B. & W. chain grates. The settings 
are installed back to back, as illustrated in Fig. 595. Each boiler has 
two 42-inch steam drums, approximately 5000 square feet of heating 
surface, and about 1000 square feet of superheating surface. Steam 
is generated at a pressure of 200 pounds per square inch with super- 
heat of 150° F. The ratio of water-heating surface to grate surface 
is approximately 55 to 1, and the ratio of the total heating surface to 
grate surface is about 66 to 1. The grates are driven by Stokes motors 
with Krehbiel oscillating engines held in reserve. The boilers are sup- 
ported by reenforced girders of the main building structure. A gallery 
is placed in front of the settings, 8 feet above the floor, to facilitate 
cleaning of tubes. Galleries are also placed between the batteries and 
on top of them. Spaces of 5 feet are provided between the sides and 
rears of the batteries, and 18 feet 8 inches in front. The furnaces are 
similar to the one illustrated in Fig. 97. The outside of the setting is 
finished with red pressed brick. Each drum is fitted with a 4|-inch pop 
safety valve. The superheater is also fitted with a pop safety valve. 
The blow-off main is 4 inches in diameter and discharges into the river. 
There are four blow-off connections to each boiler, each being provided 
with a blow-off cock and an angle valve; three of the connections are 
fitted to the mud drum and the other to the superheater drain. 

Chimneys. — One stack is provided for each section of 16 boilers. 
The shaft is supported by the steel work of the boiler setting, as shown 
in Fig. 596, an arrangement which commends itself where space is 
limited and real estate values are high. The stacks for all units are 
259 feet 6 inches in height above the grate surface, and are 18 feet in 
internal diameter. The lining is of radial fire brick and varies from 
4 inches to 13 inches in thickness. The steel sections are 5 feet high 
and vary in thickness from f inch to J inch. There are two flues, one 
32 feet long and the other 63 feet, which enter the stack on opposite 
sides. 

Turbines. — The prime movers are vertical five-stage Curtis turbines 
with base condenser and are rated at 12,000 kilowatts each. The normal 
speed is 750 r.p.m. The average steam consumption, including all 
auxiliaries, is approximately 15 pounds per kilowatt hour, corresponding 
to a coal consumption of 3 pounds per kilowatt hour (Illinois screenings, 
10,400 B.t.u. per pound). Special tests have shown as low as 12.8 
pounds per kilowatt hour, initial pressure 200 pounds gauge, 150 de- 
grees superheat, absolute back pressure J inch of mercury. Each pair 
of units has a pair of duplicate pumps, an accumulator, and a storage 
tank for supplying oil, the step-bearing pressure being maintained at 



TYPICAL CENTRAL STATIONS 



£1 




yo£ mooy Sauij 



I 



892 STEAM POWER PLANT ENGINEERING 

750 pounds per square inch. When the accumulator falls below a 
certain point a motor-driven pump is automatically started. 

Generators. — The generators are 9000-volt, 25-cycle, three-phase 
General Electric alternators mounted over the vertical shaft as illus- 
trated in Fig. 253. Exciting current is furnished by 

Two 50-kilowatt motor-driven generators. 
Two 75-kilowatt motor-driven generators. 
Two 1 50-kilowatt motor-driven generators. 
Two 75-kilowatt steam-driven generators. 
Two 1 50-kilowatt steam-driven generators. 

Part are held in reserve, though no particular units are maintained 
for the purpose. The high-tension cables lead from the generator 
through an underground tunnel to the switch house, located about 
50 feet west of the main building. The oil switches, wattmeters, and 
other instruments are located on the first floor, while the bus-bars and 
other high-tension connections are in the basement. The station switch- 
board or operating gallery in the main building is equipped with only 
such devices as are necessary for the control of the machines, all other 
instruments being located in the switch house. From the switch house 
the high-tension current is conducted through oil switches to the various 
substations, where it is converted to direct current by rotary converters, 
or transformed from 25 to 60 cycles by motor generator sets. 

The Twenty-second Street substation is located at the north end of 
the property (Fig. 593). In this substation are installed two motor 
generator sets and one rotary converter, the latter supplying direct 
current to the neighboring district and to the main station. 

Boiler and Turbine Piping. — Immediately below each boiler section 
is an apartment called the " header room," where the steam pipes from 
the various boilers join the main header, which increases in size from 
6 inches at the most remote boiler to 10 inches at the middle boiler, 
and finally to 14 inches where it leaves the nearest boiler and passes to 
the turbines. The pipes are of wrought iron, with welded flanges, and 
are packed with copper gaskets. The feeder from each boiler is 6 inches 
in diameter. An angle stop valve and a check valve are placed at the 
boiler nozzle and gate valves at the header. A motor-operated throttle 
valve and strainer are placed at the turbine, and the hydraulically oper- 
ated valves are controlled from the turbine room or the header room. 
The main header is not anchored at any point, the entire weight being 
carried by roller supports. The only drain in the header is a l|-inch 
bleeder on the boiler side of the hydraulically operated valves. The 
bleeder is connected to a. trap which discharges into either boiler or 



TYPICAL CENTRAL STATIONS 



893 




Fig. 597. General Plan of Quarry Street Station, Units 1 and 2. 



894 STEAM POWER PLANT ENGINEERING 

superheater blow-off main. All branches or feeders are drained and 
discharged into the superheater blow-off. 

Condensers and Auxiliaries. — Each unit has its own condensing appa- 
ratus, feed-water heater, hot well, and feed pumps. The condensers 
are of the Worthington "base" type with 25,000 square feet of cooling 
surface each, composed of 5900-6000 1-inch tubes 16 feet long. Cooling 
water is taken from the east slip through concrete tunnels and is dis- 
charged from the condenser into similar tunnels which empty into the 
west slip. (See Fig. 593.) 

The dry vacuum pumps are of the rotative type, with cylinders 
26 X 24, r.p.m. 100-120, and are driven by a 150-horse-power Corliss 
engine. 

The circulating pumps are of the volute single-stage centrifugal type 
and are mounted on an extension of the main shaft of the engine driving 
the dry vacuum pump. They are rated at 22,500 gallons per minute each. 

The hot-well pumps are of the two-stage centrifugal type, driven by 
20-horse-power direct-current motors. 

The feed pumps are of the Dean vertical single-cylinder pattern and 
are installed in duplicate for each unit. The steam cylinders are 24 
inches in diameter, water cylinders 14 inches in diameter, and common 
stroke 24 inches. Feed water is pumped from the hot well at a tempera- 
ture of about 100° F. and is forced through closed heaters having 3000 
square feet of heating surface, and its temperature is raised to 180 
degrees. The heater receives the steam exhausted from the steam- 
driven auxiliaries. 

From the pumps this water is forced through a 5-inch feed main to the 
different boilers in the section. The branches from main to boiler drum 
are 3 inches in diameter and fitted with an angle stop valve, a regrind- 
ing check and a gate regulating valve. A combination stop and check 
valve is placed at the drum. There is a 5-inch auxiliary main which 
supplies cold water to the boiler in case the main header is shut down. 

Miscellaneous. — The work is divided into eight-hour shifts. The 
list of operating men per unit is as follows : 

In turbine room, including janitor work 1.0 

In oil room 1.0 

Attending water 0.5 

Fireman 1.0 

Fireman's helper 0.5 

Conveyor men 1.0 

Turbine switchboard gallery 0.3 

Exciter tenders 0.2 

Switchhouse attendants 0.2 

577 



TYPICAL CENTRAL STATIONS 



895 




896 STEAM POWER PLANT ENGINEERING. 

Drains and drips from the auxiliaries empty into sumps from which 
they are discharged by Yeoman's bilge pumps into the discharge tunnel. 

A steam-driven house pump is located in the basement. 

The fire-protection system includes a 220-horse-power motor-driven 
spherical pump located in the basement and a connection for a fire tug. 

Dining room, reading room, shower and tub baths, and sleeping rooms 
for emergencies are provided for the employees. 

Quarry Street Station. — Figs. 597 to 599 give general views of the 
Quarry Street Station, which is located directly across the river from 
the Fisk Street Station. The two stations are distinct; a breakdown 
in one would not affect the other; nevertheless, they are operated to- 
gether. That is, there is one chief engineer for the two, and the com- 
bined station force of 500 men can be shifted from one to the other as 
needed. 

The general layout of the Quarry Street Station differs from that of 
the Fisk Street Station on account of real-estate limitations. The 
boilers are in two parallel rows instead of the equipment for each unit 
extending at right angles to the turbine room as at Fisk. 

The station contains six 14,000-kilowatt Curtis turbo-alternators, the 
steam for each unit being supplied by eight 500-horse-power B. & W. 
boilers arranged as shown in Fig. 597. A 5000-kilowatt motor-generator 
set is also included in the equipment. Steam is generated at 225 pounds 
gauge pressure and 150-175 degrees superheat. The settings for the 
first two units are similar to that illustrated in Fig. 99, and the other 
similar to the one illustrated in Fig. 100. 

A novel system of ventilation enables the generators to be operated 
continuously at full load. As will be seen from Fig. 598 air ducts lead 
from an outside intake to the top of each unit, the revolving portion of 
the generator being designed to draw in a continuous supply of air and 
discharge it through openings in the turbine casing. 

Everything is in duplicate, so that the chance of breakdown is remote. 
Each unit is equipped with a volute centrifugal pump driven by a 125- 
horse-power Corliss engine. The water of condensation is removed 
from the condensers by hot-well pumps driven by Kerr steam turbines, 
one for each unit. For each two units of turbines and boilers three 
boiler feed pumps are provided, two centrifugal and one reciprocating, 
all of them located between the turbines. There are also four step- 
bearing oil pumps, two oil accumulators, dry-air pumps, oil filters, etc. 
All are in plain view of the turbine-room operating force. 

For the six units there are five 150-kilowatt exciters, three driven by 
horizontal Curtis steam turbines and two by induction motors. In 
addition there is an excitation storage battery of 70 cells in the base- 



TYPICAL CENTRAL STATIONS 



897 




Fig. 599. Sectional Elevation, Quarry Street Station. 



898 



STEAM POWER PLANT ENGINEERING 



ment. Furthermore, in an emergency, 500-kilowatt rotary converters 
of the substation could be used for excitation. 

Northwest Stations. — Two new stations of 120,000-kilowatt rated 
capacity each are to be installed on the north branch of the Chicago 
River near Roscoe Street and California Avenue. Each station is to be 
equipped with six 20,000-kilowatt Curtis turbo-generators, 2300-volt, 
25-cycle, three-phase, 750-r.p.m., similar to those installed at Quarry 
Street. The first two units for one station are now in operation. Each 
unit is supplied with steam at 250 pounds gauge pressure and 150 
degrees superheat from ten 500-horse-power B. & W. boilers. The 
boiler settings are similar to the one illustrated in Fig. 100. The 
general layout is the same as at Fisk, the boiler lanes extending at 
right angles to the turbine room- There is one chimney 17 feet inside 
diameter and 250 feet in height for every ten boilers. 

The present capacity (November, 1912) of the Commonwealth Edi- 
son Company is about 280,000 kilowatts, divided as follows: 





Units. 


Present 
Capacity. 


Ultimate 
Capacity. 


Fisk 

Quarry 

Northwest No. 1 

Northwest No. 2 


10-12,000 
6-14,000 
6-20,000 
6-20,000 


120,000 
84,000 
40,000 


170,000* 

84,000 

120,000 

120,000 

36,000 


Miscellaneous Plants 


36,000 
280,000 






530,000 



2-25,000-kilowatt Parsons turbines are now being added to the present equipment. 
COMPARATIVE BOILER ROOM AND ENGINE ROOM AREAS. 





Fisk. 


Quarry. 


Northwest. 


Boiler room, sq. ft. per kw 

Engine room, sq. ft. per kw 

Total area, sq. ft. per kw 


0.51 

0.24 
0.75 




0.44 


0.18 


0.15 
0.59 







For a detailed description of the Northwest Station, consult Practical Engineer, U. S., Jan. 15, 
1913,'p. i09. 



CHAPTER XXI. 

A TYPICAL MODERN ISOLATED STATION.* 
Bleeder Turbines and Condenser System. 

The new power plant of the W. H. McElwain Company at Man- 
chester, N. H., is an excellent example of current practice in generation 
of power by steam for industrial purposes. 

General Arrangement. — General arrangement of the boiler and 
engine rooms is shown in plan in Fig. 601. At the present time there 
have been installed three 300-horse-power water-tube boilers and one 
1000-kilowatt turbo-generator outfit. The boiler room contains suffi- 
cient space for a fourth 300-horse-power unit, as indicated by dotted 
lines. The completed plant will include duplicates of the two batteries 
shown, making a total of 2400 horse power. The future boilers will 
face those already installed, the building being extended for this pur- 
pose, and the firing space shown will be common to both sections. 

The chimney, which is 176 feet in height, with a flue 9 feet in diam- 
eter, is designed with reference to the final capacity of the plant. In 
the engine room, at the right, is shown space for two additional generating 
units, which provide for an ultimate capacity of 3000 kilowatts. Sec- 
tional elevations, showing the boilers, turbines, and the various auxiliary 
equipment and their connections, are illustrated in Figs. 600, 602, and 603. 

Boilers. — Present equipment consists of three Babcock and Wilcox 
water-tube boilers, each containing 2972 square feet of heating surface 
and about 50 square feet of grate surface. The heating surface is made 
up of two steam drums, tubes, and mud drum, and a superheater of the 
form shown in Fig. 602. 

Each boiler contains 144 4-inch tubes, 18 feet in length, made up in 
12 sections of 12 tubes each, and 2 steam drums, 3 feet in diameter by 
20 feet 4 inches in length. The superheaters each contain approxi- 
mately 372 square feet of surface, which is 12J per cent of the heating 
surface of the boiler, and are designed to give 100 degrees superheat 
when the boilers are operated at their normal rating of 300 horse power. 
The proportions of all parts are designed for a working pressure of 160 
pounds per square inch and the safety valves are set at that point. 

* From the Practical Engineer, Chicago, July 1, 1912. 



900 



STEAM POWER PLANT ENGINEERING 




A TYPICAL MODERN ISOLATED STATION 



901 




902 STEAM POWER PLANT ENGINEERING 

Each boiler is provided with a water column fitted with high and low 
water alarm, try cocks, and gauge glass with special device for shutting 
off in case of breakage. Also 3J-inch lock pop-safety valve, and 12-inch 
steam gauge reading to 300 pounds pressure. The feed pipes are 2 
inches in diameter, provided with both check and gate valves, the latter 
having special extension handles. The blow-off connections are of 
2J-inch extra heavy pipe, and are each provided with two blow-off 
valves of special design. 

Boiler settings are of hard-burned brick, laid in cement mortar, 
consisting of 1 part cement to 3 of sand, up to the level of the grates, 
and in lime mortar above that point. All parts of the furnaces and 
setting exposed to the fire are lined with firebrick laid in fire clay. The 
furnaces are of the "Dutch oven" type as shown in Fig. 602. 

Smoke Connections. — Location of the main smoke flue is best 
shown in Fig. 601. It is 4 feet 9 inches by 7 feet 6 inches in size and 
constructed of No. 10 black iron. It is stiffened with angle-iron 
braces and supported from the roof. The uptake from each boiler 
is provided with an adjusting damper for hand manipulation from 
the floor level. 

A balanced damper is located in the main flue at the point indicated, 
and operated by an automatic regulator of the hydraulic type. An 
interesting detail in connection with this work is the method of attach- 
ing the covering to the lower side of the flue so that it will not sag or 
peel off. This consists of cross pieces of 1-inch tee-bars placed 24 inches 
apart and riveted to the flue. The projecting flanges of these bars are 
drilled at frequent intervals and wires strung through, to which the 
covering is attached. 

Handling of Fuel and Ash. — Coal is brought to the fire room by 
cars running on a special track as shown in Fig. 601. This track passes 
over platform scales just inside the building, where each load may be 
weighed as it is brought in. The track is double within the fire room 
so that the cars may pass, and also to furnish storage space for both 
coal and ash cars when so desired. 

The arrangement for the removal of ash is best shown in Figs. 601 
and 602. A dumping chute is provided in the bottom of each ashpit 
and at such an elevation that a car may be run underneath it as indi- 
cated. When filled, they are pushed to the ash lift (see Fig. 601) where 
they are raised to the boiler-room level and run out on the coal track 
for disposal. Combustible waste from the factory is brought through 
a 36-inch pipe to a collector placed in the upper part of the boiler 
room, as shown in Fig. 602, and fed into the furnaces as there 
indicated. 



A TYPICAL MODERN ISOLATED STATION 



903 



Turbine and Generator. — The turbo-generator unit is one of the 
Westinghouse make, of 1000-kilowatt capacity, and equipped with an 
automatic bleeder connection and constant-pressure valve. It is 6 feet 




6 inches in width by 24 feet 8 inches in length and weighs approxi- 
mately 79,000 pounds. It is of the regular Westinghouse-Parsons type, 
the most interesting feature being the bleeder attachment which adapts 
it for use in combined power and heating plants. An important re- 
quirement for the economical operation of the ordinary steam turbine 



904 



STEAM POWER PLANT ENGINEERING 



is the maintenance of a high vacuum at the exhaust end, which, of 
course, prevents the utilization of exhaust steam for heating purposes. 

The capacity of the turbine under different conditions is as follows: 
With a throttle pressure of 150 pounds per square inch (gauge), a vacuum 
of 28 inches, 100 degrees superheat, and a speed of 3600 r.p.m., the 
normal capacity when condensing is 1500 b.h.p. and the maximum 
2250 b.h.p. When running non-condensing with a back pressure not 
exceeding that of the atmosphere, the maximum capacity is 1500 b.h.p. 

It is interesting to note the probable steam economy of a turbine of 
this type when operating under varying loads, as expressed in the 
guarantee placed upon this machine, which is as follows: When operat- 
ing under the above conditions, in connection with the generator 
attached, the steam consumption per hour, including all leakage and 
loss with the turbine, shall not exceed the quantities given below: 



Load, 


Power Factor, 


Kilowatts. 


Pounds Steam per 


Per Cent. 


Per Cent. 


Kilowatt-Hour. 


150 


80 


1500 


18.8 


125 


80 


1250 


18.3 


100 


80 


1000 


17.9 


75 


80 


750 


18.8 


50 


80 


500 


20.7 



When operating under the same general conditions, with 3 pounds 
gauge pressure at the bleeder connection, the steam consumption 
per hour shall not exceed the following at the loads indicated, when 
withdrawing the following amounts of steam through the bleeder 
connection : 







Pounds of Steam 


Steam to Condenser. 


Load, 


Kilowatts. 






Per Cent. 














To Throttle. 


To Bleeder. 


Total. 


Kilowatts. 


150 


1500 


38,000 


18,600 


19,400 


12.9 






31,000 


10,000 


21,000 


14.0 


125 


1265 


38,000 


22,000 


16,000 


12.7 






36,300 


20,000 


16,300 


12.9 






29,200 


10,000 


19,200 


15.2 


100 


1000 


37,200 


30,000 


7,200 


7.2 






30,000 


20,000 


10,000 


10.0 






24,400 


10,000 


14,400 


14.4 


75 


716 


30,500 


30,000 


500 


0.7 






25,600 


20,000 


5,600 


7.8 






20,200 


10,000 


10,200 


14.2 


50 


469 


21,700 


21,700 





0.0 






20,600 


20,000 


600 


1.3 






16,000 


10,000 


6,000 


12.8 



A TYPICAL MODERN ISOLATED STATION 



905 



Generator. — The generator is of the revolving-field type with 
inclosed frame, generating a 3-phase, 60-cycle, alternating current of 
600 volts. The efficiency rating, with a power factor of 100 per cent, 
is as follows : 



Load, 
Per Cent. 


Efficiency, 
Per Cent. 


Load, 
Per Cent. 


Efficiency, 
Per Cent. 


50 

75 

100 


90.10 
93.00 

94.50 


125 

150 


95.50 

95.75 



Temperature rise based on its normal rating and a power factor of 
80 per cent, for periods of different length and for various loads, is given 
below: 



Load, 
Per Cent. 


Length of Run, 
Hours. 


Temperature Rise, 
Armature. 


Degs. F., Field. 


100 
125 
150 


24 

24 

1 


72 

90 

108 


72 

90 

108 



The maximum conditions of continuous operation with a power factor 
of 80 per cent and for a room temperature of 77 degrees F. are as follows: 
Output, 1250 kilowatts (25 per cent overload). Rise in temperature: 

Armature, 90 degrees F. 
Field, 90 degrees F. 

Maximum temperature to which insulation can be subjected without 
injury: 

Armature, 194 degrees F. 
Field, 302 degrees F. 

There are two exciters provided, one being turbine driven and having 
a normal capacity of 25 kilowatts; the other motor driven, with a 
capacity of 40 kilowatts. The turbine is of the Westinghouse make, 
horizontal type, with a normal capacity of 38 b.h.p. at a speed of 3500 
r.p.m. when running non-condensing, and a continuous overload capac- 
ity of 25 per cent. The steam requirements for this machine as re- 
gards temperature and pressure are the same as for the main turbine. 

The exciter is a direct-current machine with shunt winding, generat- 
ing a current of 125 volts at full load. 

Condensing Apparatus. — In connection with the main turbine a 
Westinghouse-LeBlanc jet condenser is used, and is shown in elevation 
in Figs. 602 and 603. This is designed to operate under a normal lift 



906 



STEAM POWER PLANT ENGINEERING 




A TYPICAL MODERN ISOLATED STATION 



907 



of 18 feet and takes its water supply from the intake tunnel as shown. 
When using injection water at a temperature of 70 degrees F. the fol- 
lowing results are guaranteed, with a water consumption not exceeding 
724,000 pounds per hour: 



Steam Condensed 

per Hour, 

Pounds. 


Vacuum Main- 
tained, Inches 
(Barometer, 30 Ins.) 


Steam Condensed 
per Hour, 
Pounds. 


Vacuum Main- 
tained, Inches 
(Barometer, 30 Ins.) 


10,350 
14,100 
18,000 


28.65 

28.44 
28.17 


19,950 
22,900 
30,000 


28.00 
27.80 
27.11 



The vacuum air pump is of the turbine type and is mounted upon 
the same shaft with the centrifugal ejector pump, both being driven by 
a steam turbine of 41 b.h.p. running at 1500 r.p.m. under an atmos- 
pheric exhaust pressure. This piece of apparatus is shown at the base 
of the condenser in Figs. 602 and 603. 

High-pressure Piping System. — This includes all high-pressure piping 
in the boiler and engine rooms for the supply of turbines, pumps, etc., 
and for the supplementary supply to the heating system as may be 
needed. Pipe used for this purpose is full weight, wrought iron being 
used for sizes below 6 inches and open-hearth steel for larger sizes. 
The main drum at the rear of the boilers is of gun metal with nozzles 
cast in place. Expansion is provided for, so far as possible, by the use 
of sweep pipe bends and fittings of the long-turn pattern, all 2J-inch 
and larger fittings being of this design with flange joints. The high- 
pressure connections are shown in Figs. 601, 602, and 604. Starting 
at the boilers (Fig. 602), 6-inch leads are carried to a 12-inch drum 
supported on low piers and rolls at the rear of the boilers. From here 
a 6-inch branch leads to the main turbine, and two branches of the same 
size to a 6-inch auxiliary main, running beneath the engine-room floor, 
near to, and parallel with, the boiler-room wall. From this auxiliary 
main are taken the supplies to the various minor turbines and pumps, 
and also the branches leading to the low and intermediate-pressure 
system through reducing valves. The main drum is divided into two 
sections by means of a valve at the center, and each of these sections 
is connected with the auxiliary drum as shown in Figs. 600 and 604. 
The supplies to the various pumps are easily traced from Fig. 604, also 
the connections with the 18-inch heating main and the intermediate- 
pressure line, leading to the factory through the tunnel leaving the 
building as indicated in the upper right-hand corner of the drawing. 

Exhaust System. — All low and intermediate pressure piping is full 
weight, sizes up to, and including, 12-inch being of wrought iron, while 



908 



STEAM POWER PLANT ENGINEERING 




A TYPICAL MODERN ISOLATED STATION 909 

open-hearth steel is employed for the larger sizes. Standard-weight fit- 
tings are used for this work, those 6 inches and over being of the long- 
turn pattern. Flange joints are provided on all piping 2| inches and 
larger in diameter, the same as for high-pressure work. The exhaust 
piping is most clearly shown in Figs. 602, 603, and 604. Referring to 
Fig. 604 the 18-inch exhaust from the main turbine is shown as leading 
through a back-pressure valve into a 30-inch outboard line designed for 
the completed plant. This is clearly shown in elevation in Fig. 603. 
An 8-inch auxiliary exhaust connecting with the various pumps is 
shown in Fig. 603, parallel with, and below, the auxiliary high-pressure 
main already described. Steam from this enters the heating system 
through an oil separator. The 12-inch bleeder connection from the 
turbine leads to the 18-inch heating main and is shown in the same 
drawing, although more clearly in Fig. 604. 

Drainage. — The blow-off main from the boilers is carried directly to 
the river through a 4-inch cast-iron pipe. Drips from high-pressure 
piping are trapped to the main receiving tank and pumped back to the 
boilers. Exhaust drips, and all condensation containing oil, are trapped 
to a cast-iron sump tank located in the condenser pit, and, together 
with other drainage, are discharged by means of a water ejector. 

Water Supply, Feed Piping, etc. — Water for condensing and fire 
purposes is brought from the river through a cement conduit, a section 
of this, together with the 15-inch suction to the condenser, being shown 
in Figs. 602 and 603. The discharge from the condenser pump is into 
an 18-inch pipe leading to the river and shown in section in Fig. 602. 
Water pressure for fire protection is furnished by an 18 by 10 by 12-inch 
Underwriters' fire pump of 1000 gallons capacity, placed in the con- 
denser pit; this is shown in elevation in Fig. 603 and in plan in Fig. 604 
and takes its supply from the intake tunnel as there shown. 

The house tank and boilers have two sources of supply, one directly 
from the city mains and the other from the intake tunnel. There is 
also a tank arrangement whereby water may be drawn from the dis- 
charge pipe of the condenser pump. ' 

These various lines are shown in Fig. 605. A 6-inch connection from 
the city main enters as shown at the upper part of the drawing, toward 
the left, and, after passing through a meter, branches are carried to the 
house tank, the receiving tank, the boilers, and to the priming pipes of 
the condenser and vacuum pumps. 

The second source of supply, that from the intake tunnel, requires 
the use of two turbine-driven house pumps of the one-stage turbine 
type, located in the condenser pit as shown in Fig. 605. These 
pumps each have a capacity of 200 gallons per minute against a head 



910 



STEAM POWER PLANT ENGINEERING 



of 150 feet, and discharge into a line of piping having branches con- 
necting with the house tank, receiving tank, and boiler feed pipe. A 
filtering equipment is also provided, as shown in Fig. 605, and so con- 
nected that the water from this source may be purified if desired. 

Boiler Feed. — Feed lines connecting with the boilers are shown in 
Fig. 601. One of these supplies water either directly by city pressure 



7/////////////// /7777777MP/. 




6 Main Return , 



3" Suction Strainer ' 



Fig. 605. Plan of Condenser Piping. 



or from the turbine house pumps. The other supply is from a pair 
of boiler feed pumps connecting with a receiving tank located in the 
boiler room as shown. The feed pumps, two in number, are of the 
duplex, outside packed, pot-valve type, 8 by 5 by 10 inches in size. 
The tank is 4 feet in diameter by 6 feet in length, of f-inch iron plate, 
and is connected with both city pressure and the house pumps. Under 
ordinary working conditions the feed supply is first discharged into the 



A TYPICAL MODERN ISOLATED STATION 911 

tank and then pumped to the boilers through a heater of 1000-horse- 
power capacity located as shown in the drawing. 

Heating System. — Factory buildings are heated by direct radiation 
in the form of coils and cast-iron radiators as best suited to local con- 
ditions. The Webster system of circulation is employed, a pair of 
6 by 10 by 12-inch single-piston vacuum pumps being connected with 
the main return as shown in Fig. 605. These discharge into the receiv- 
ing tank in the boiler room, and the condensation is pumped back to 
the boilers with the fresh feed. 

Steam supply for the radiation has already been mentioned, coming 
principally through the bleeder connection from the main turbine, 
supplemented, when necessary, by live steam through a reducing valve. 

Insulation. — In general, tanks, heaters, etc., are covered with 85 
per cent magnesia blocks, finished with a plastic coat of the same 
material, the total thickness of the covering, when finished, ' being 2 
inches. In addition to this, tanks and heater are provided with a 
covering of 7-ounce canvas. The insulation on that portion of the 
smoke pipe which comes outside of the building is protected by a cover- 
ing of heavy sheet iron. Steam piping, both high and low pressure, 
is insulated with 85 per cent magnesia sectional covering. All cold- 
water piping, with the exception of the connections to the condenser, are 
covered with wool felt, having a lining of tarred paper. Pipe covering 
of all kinds is finished with a heavy canvas jacket and painted. 



APPENDIX A. 

INSTRUCTIONS REGARDING TESTS IN GENERAL.* 
A.S.M.E. CODE OF 1912. 

1. OBJECT. 

Ascertain the specific object of the test, and keep this in view not 
only in the work of preparation, but also during the progress of the test, 
and do not let it be obscured by devoting too close attention to matters 
of minor importance. Whatever the object of the test may be, accuracy 
and reliability must underlie the work from beginning to end. 

If questions of fulfillment of contract are involved, there should be 
a clear understanding between all the parties, preferably in writing, as 
to the operating conditions which should obtain during the trial, and 
as to the methods of testing to be followed, unless these are already 
expressed in the contract itself. 

Among the many objects of performance tests, the following may be noted: 

Determination of capacity and efficiency, and how these compare with 
standard or guaranteed results. 

Comparison of different conditions or methods \of operation. 

Determination of the cause of either inferior or superior results. 

Comparison of different kinds of fuel. 

Determination of the effect of changes of design or proportion upon capacity 
or efficiency, etc. 

2. PREPARATIONS. 

(A) Dimensions. 

Measure the dimensions of the principal parts of the apparatus to 
be tested, so far as they bear on the objects in view, or determine these 
from correct working drawings. Notice the general features of the 
same, both exterior and interior, and make sketches, if needed, to show 
unusual points of design. 

(B) Examination of Plant. 

Make a thorough examination of the physical condition of all parts 
of the plant or apparatus which concern the object in view, and record 
the conditions found, together with any points in the matter of opera- 
tion which bear thereon. 

* Preliminary report of the Committee on Power Tests (Jour. A.S.M.E., Nov., 
1912). Greatly abridged. 

912 



APPENDIX A 913 

If the object of the test is to determine the highest efficiency or 
capacity obtainable, any physical defects, or defects of operation, 
tending to make the result unfavorable should first be remedied; all 
fouled parts being cleaned, and the whole put in first-class condition. 
If, on the other hand, the object is to ascertain the performance under 
existing conditions, no such preparation is either required or desired. 

(C) General Precautions against Leakage. 

In steam tests make sure that there is no leakage through blow-offs, 
drips, etc., or any steam or water connections of the plant or apparatus 
undergoing test, which would in any way affect the results. All such 
connections should be blanked off, or satisfactory assurance should be 
obtained that there is leakage neither out nor in. This is a most 
important matter, and no assurance should be considered satisfactory 
unless it is susceptible of absolute demonstration. 

(D) Apparatus and Instruments. 

Select the apparatus and instruments specified in the Code of Rules 
applying to the test in hand, locate and install the same, and complete 
the preparations for the work in view. 

The arrangement and location of the testing appliances in every case 
must be left to the judgment and ingenuity of the engineer in charge, 
the details being largely dependent upon locality and surroundings. 
One guiding rule, however, should always be kept in view, viz., see that 
the apparatus and instruments are substantially reliable, and arrange 
them in such a way as to obtain correct data. 

3. MISCELLANEOUS INSTRUCTIONS. 

The person in charge of a test should have the aid of a sufficient 
number of assistants, so that he majr be free to give special attention 
to any part of the work whenever and wherever it may be required. 
He should make sure that the instruments and testing apparatus 
continually give reliable indications, and that the readings are correctly 
recorded. He should also keep in view, at all points, the operation 
of the plant or part of the plaut under test and see that the operating 
conditions determined on are maintained and that nothing occurs, either 
by accident or design, to vitiate the data. This last precaution is 
especially needed in guarantee tests. 

Before a test is undertaken, it is important that the boiler, engine, 
or other apparatus concerned, shall have been in operation a sufficient 
length of time to attain working temperatures and proper operating 
conditions throughout, so that the results of the test may express the 
true working performance. 



914 STEAM POWER PLANT ENGINEERING 

An exception should be noted where the object of the test is to obtain 
the working performance, including the effect of preliminary heating, 
in which case all the conditions should conform to those of regular 
service. 

In preparation for a test to demonstrate maximum efficiency, it is 
desirable to run preliminary tests for the purpose of determining the 
most advantageous conditions. 

4. OPERATING CONDITIONS. 

In all tests in which the object is to determine the performance under 
conditions of maximum efficiency, or where it is desired to ascertain 
the effect of predetermined conditions of operation, all such conditions 
which have an appreciable effect upon the efficiency should be main- 
tained as nearly uniform during the trial as the limitations of practical 
work will permit. In a stationary steam plant, for example, where 
maximum efficiency is the object in view, there should be uniformity 
in such matters as steam pressure, times of firing, quantity of coal 
supplied at each firing, thickness of fire, and in other firing operations; 
also in the rate of supplying the feed water, in the load on the engine 
or turbine, and in the operating conditions throughout. On the other 
hand, if the object of the test is to determine the performance under 
working conditions, no attempt at uniformity is either desired or re- 
quired unless this uniformity corresponds to the regular practice, and 
when this is the object the usual working conditions should prevail 

throughout the trial. ' 

5. RECORDS. 

A log of the data should be entered in notebooks or on blank sheets 
suitably prepared in advance. This should be done in such manner 
that the test may be divided into hourly periods, or, if necessary, periods 
of less duration, and the leading data obtained for any one or more 
periods as desired, thereby showing the degree of uniformity obtained. 

The readings of the various instruments and apparatus concerned 
in the test, other than those showing quantities of consumption (such 
as fuel, water, and gas), should be taken at intervals not exceeding half 
an hour and entered in the log. Whenever the indications fluctuate, 
the intervals should be reduced according to the extent of the fluctua- 
tion. In the case of smoke observations, for example, it is often neces- 
sary to take observations every minute, or still oftener, continuing 
these throughout the period covering the range of variations. 

Make a memorandum of every event connected with the progress 
of a test, however unnecessary at the time it may appear. A record 
should be made of the exact time of every such occurrence and the 



APPENDIX A 915 

time of taking every weight and every observation. For the purpose 
of identification the signature of the observer and the date should be 
affixed to each log sheet or record. 

In the simple matter of weighing coal by the barrow-load, or weighing 
water by the tank-full, which is required in many tests, a series of marks, 
or tallies, should never be trusted. The time each load is weighed or 
emptied should be recorded. The weighing of coal should not be 
delegated to unreliable assistants, and, whenever practicable, one or 
more men should be assigned solely to this work. The same may be 
said with regard to the weighing of feed water. 

6. DATA REPRESENTED GRAPHICALLY. 

If it is desired to show the uniformity of the data at a glance the 
whole log of the trial should be plotted on a chart, using horizontal 
distances to represent times of observation, and vertical distances on 
suitable scales to represent various data as recorded. Such a chart 
showing log of a boiler test is illustrated in Appendix No. 23. 

It is instructive to plot the leading data on such a chart while the 
test is in progress. 

7. REPORT. 

The report of a test should present all the leading facts bearing 
on the design, dimensions, condition, and operation of the apparatus 
tested, and should include a description of any other apparatus and 
auxiliaries concerned, together with such sketches as may be needed 
for a clear understanding of all points under consideration. It should 
state clearly the object and character of the test, the methods followed, 
the conditions maintained, and the conclusions reached, closing with 
a tabular summary of the principal data and results. 



APPENDIX B. 

RULES FOR CONDUCTING EVAPORATIVE TESTS OF BOILERS.* 
A.S.M.E. Code of 1912. 

1. OBJECT AND PREPARATIONS. 

Determine the object, take the dimensions, note the physical con- 
ditions, examine for leakages, install the testing appliances, etc., as 
pointed out in the general instructions and make preparations for the 

test accordingly. 

2. FUEL. 

Determine the character of fuel to be used. For tests of maximum 
efficiency or capacity of the boiler to compare with other boilers, the 
coal should be of some kind which is commercially regarded as a standard 
for the locality where the test is made. 

A coal selected for maximum efficiency and capacity tests should be 
the best of its class, and especially free from slagging and unusual 
clinker-forming impurities. 

For guarantee and other tests with a specified coal containing not 
more than a certain amount of ash and moisture, the coal selected 
should not be higher in ash and in moisture than the stated amounts, 
because any increase is liable to reduce the efficiency and capacity more 
than the equivalent proportion of such increase. 

The size of the coal, especially where it is of the anthracite class, 
should be determined by screening a suitable sample. 

3. APPARATUS AND INSTRUMENTS. 
The apparatus and instruments required for boiler tests are : 

(a) Platform scales for weighing coal and ashes. 

(b) Graduated scales attached to the water glasses. 

(c) Tanks and platform scales for weighing water (or water meters calibrated in 

place). 

(d) Pressure gages, thermometers, and draft gages. 

(e) Calorimeters for determining the calorific value of fuel and the quality of 

steam. 
(/) Furnace pyrometers. 
(g) Gas analyzing apparatus. 

* Preliminary report of the Committee for Power Tests. (Jour. A.S.M.E., 
Nov., 1912.) Greatly abridged. 

916 



APPENDIX B 917 

Full directions regarding the use and calibration of the above men- 
tioned appliances are given in the complete code. 

4. OPERATING CONDITIONS. 

Determine what the operating conditions and method of firing should 
be to conform to the object in view, and see that they prevail through- 
out the trial, as nearly as possible. 

5. DURATION. 

The duration of tests to determine the efficiency of a hand-fired 
boiler should be 10 hours of continuous running, or such time as may 
be required to burn a total of 250 pounds of coal per square foot of grate. 

In the case of a boiler using a mechanical stoker, the duration, where 
practicable, should be at least 24 hours. If the stoker is of a type that 
permits the quantity and condition of the fuel bed at beginning and 
end of the test to be accurately estimated, the duration may be reduced 
to 10 hours, or such time as may be required to burn the above-noted 
total of 250 pounds per square foot. 

6. STARTING AND STOPPING. 

The conditions regarding the temperature of the furnace and boiler, 
the quantity and quality of the live coal and ash on the grates, the water 
level, and the steam pressure, should be as nearly as possible the same 
at the] end as at the beginning of the test. 

7. RECORDS. 

The records of data should be obtained as pointed out in Appendix 
A. Half -hourly readings of the instruments are usually sufficient. If 
there are sudden and wide fluctuations, the readings in such cases should 
be taken every fifteen minutes, and in some instances oftener. 

8. QUALITY OF STEAM. 

If the boiler does not produce superheated steam the percentage of 
moisture in the steam should be determined by the use of a throttling 
or separating calorimeter. If the boiler has superheating surface, the 
temperature of the steam should be determined by the use of a ther- 
mometer inserted in a thermometer well. 

9. SAMPLING AND DRYING COAL. 

During the progress of the test the coal should be regularly sampled 
for the purpose of analysis and determination of moisture. 



918 STEAM POWER PLANT ENGINEERING 

10. ASHES AND REFUSE. 

The ashes and refuse withdrawn from the furnace and ashpit during 
the progress of the test and at its close should be weighed so far as 
possible in a dry state. If wet the amount of moisture should be 
ascertained and allowed for, a sample being taken and dried for this 
purpose. This sample may serve also for analysis and the determina- 
tion of unburned carbon and fusing temperature. 

11. CALORIFIC TESTS AND ANALYSES OF COAL. 

The quality of the fuel should be determined by calorific tests and 
analysis of the coal sample above referred to. 

12. ANALYSES OF FLUE GASES. 

For approximate determinations of the composition of the flue gases, 
the Orsat apparatus, or some modification thereof, should be employed. 
If momentary samples are obtained the analyses should be made as 
frequently as possible, say every 15 to 30 minutes, depending on the 
skill of the operator, noting at the time the sample is drawn, the furnace 
and firing conditions. If the sample drawn is a continuous one, the 
intervals may be made longer. 

13. SMOKE OBSERVATIONS. 

In tests of bituminous coals requiring a determination of the amount 
of smoke produced, observations should be made regularly throughout 
the trial at intervals of five minutes (or if necessary every minute), 
noting at the same time the furnace and firing conditions. 

14. CALCULATION OF RESULTS. 

The methods to be followed in expressing and calculating those results 
which are not self-evident are explained as follows: 

(a) Efficiency. The " efficiency of boiler, furnace, and grate" is the relation 
between the heat absorbed per pound of coal fired and the calorific value of 
one pound of coal. 

The "efficiency of boiler and furnace" is the relation between the heat 
absorbed per pound of combustible burned, and the calorific value of one 
pound of combustible. This expression of efficiency furnishes a means for 
comparing one boiler and furnace with another, when the losses of unburned 
coal due to grates, cleanings, etc., are eliminated. 

The "combustible burned" is determined by subtracting from the weight 
of coal supplied to the boiler, the moisture in the coal, the weight of ash and 
unburned coal withdrawn from the furnace and ashpit, and the weight of 
dust, soot, and refuse, if any, withdrawn from the tubes, flues, and combustion 



APPENDIX B 919 

chambers, including ash carried away in the gases, if any, determined from 
the analyses of coal and ash. The "combustible" used for determining the 
calorific value is the weight of coal less the moisture and ash found by analysis. 
The "heat absorbed" per pound of coal, or combustible, is calculated by 
multiplying the equivalent evaporation from and at 212 degrees per pound 
of coal or combustible by 970.4. 

(b) Corrections for Moisture in Steam. When the percentage is less than 2 per 
cent it is sufficient merely to deduct the percentage from the weight of water 
fed. If the percentage is greater than 2 per cent or if extreme accuracy is 
required, the factor of correction equals 

V + r {H-h) 1 

in which Q is the quality of the steam (one minus the decimal representing 
the percentage of moisture), P the proportion of moisture, T the total heat 
of water at the temperature of the steam, h the total heat of the feed water, 
and H the total heat of saturated steam of the given temperature. 

(c) Correction for Live Steam, if Any, used for Aiding Combustion. If live steam 
is admitted into the furnace or ashpit for producing blast, injecting fuel, or 
aiding combustion, it is to be deducted from the total evaporation, and the 
net evaporation used in the various calculations. 

(d) Equivalent Evaporation. The equivalent evaporation from and at 212 de- 
grees is obtained by multiplying the weight of water evaporated, corrected for 
moisture in steam, by the "factor of evaporation." The latter equals 

H-h 
970.4 ' 

in which H and h are respectively the total heat of saturated steam and of the 
feed water entering the boiler. When the steam is superheated, the total heat 
of the steam is that of saturated steam plus the product of the number of 
degrees of superheating by the specific heat of the steam. 

Unless otherwise provided, a combined boiler and superheater should be 
treated as one unit, and the equivalent of the work done by the superheater 
should be included in the evaporative work of the boiler. 

(e) Heat Balance. The "heat balance," or approximate distribution of the 
calorific value of the coal or combustible among the several items of heat 
utilized and heat lost, should be obtained in cases where the flue gases have 
been analyzed and a complete analysis made of the coal. 

The loss due to moisture in the coal is found by multiplying the total 
heat of one pound of superheated steam at the temperature of the escaping 
gases, calculated from the temperature of the air in the boiler room, by the 
proportion of moisture. 

The loss due to moisture formed by the burning of hydrogen is obtained by 
multiplying the total heat of one pound of superheated steam at the tem- 
perature of the escaping gases, calculated from the temperature of the air in 
the boiler room, by the proportion of the hydrogen, determined from the 
analysis of the coal, and multiplying the result by 9. 

The loss due to heat carried away in the dry gases is found by multiplying 
the weight of gas per pound of coal or combustible by the elevation of tem- 
perature of the gases above the temperature of the boiler room, and by the 



920 STEAM POWER PLANT ENGINEERING 

specific heat of the gases (0.24). The weight of gas referred to is obtained by- 
finding the weight of dry gas per pound of carbon burned, using the formula 

11C0 2 + 8Q + 7(C0 + N) 
3 (C0 2 + CO) 

in which CO2, CO, O, and N are expressed in percentages by volume, and 
multiplying this result by the proportion borne by the carbon burned to the 
whole amount of coal or combustible as determined from the results of the 
analysis of the coal, ash, and refuse. 

The loss due to incomplete combustion of carbon is found by first obtaining 
the proportion borne by the carbon monoxide in the gases to the sum of the 
carbon monoxide and carbon dioxide, and then multiplying this proportion 
by the proportion of carbon in the coal or combustible, and finally multiplying 
the product by 10,150, which is the number of heat units generated by burn- 
ing to carbon dioxide one pound of carbon contained in carbon monoxide. 

The loss due to combustible matter in the ash and refuse is found by mul- 
tiplying the proportion that this combustible bears to the whole amount of 
coal or combustible, by its calorific value per pound. For most purposes it 
is sufficient to assume the latter to be 14,600 B.t.u., the same as that of carbon. 

The loss due to moisture in the air is determined by multiplying the weight 
of such moisture per pound of coal or combustible by the elevation of tem- 
perature of the flue gases above the temperature of the boiler room and by 
0.47. The weight of moisture is found by multiplying the weight of air per 
pound of coal or combustible by the moisture in one pound of air determined 
from readings of the wet and dry-bulb thermometer. 
(/) Total Heat of Combustion of Coal, by Analysis. The total heat of combustion 
may be computed from the results of the ultimate analysis by using the formula 



14,600 C + 62,000 [ H - ^ ) + 4000 S 



(-8) 



in which C, H, O, and S refer to the proportions of carbon, hydrogen, oxygen, 

and sulphur, respectively. 

(g) Air for Combustion. The quantity of air used may be calculated by the 

formulae : 

tu t • iu r u 3.032N 

Lb. of air per lb. of carbon = ~^c — ~^ > 

in which N, C0 2 , and CO are the percentages of dry gas obtained by analysis, 

and 

Lb. of air per lb. of coal = lb. air per lb. C X per cent C in the coal. 

The ratio of the air supply to that theoretically required for complete 

N 



combustion is 



N - 3.782 O 



15. DATA AND RESULTS. 

The data and results should be reported in accordance with either 
the Short Form or the Complete Form printed below, adding lines for 
data not provided for, or omitting those not required, as may conform 
to the object in view. 



APPENDIX B 



921 



16. CHART. 

In trials having for an object the determination and exposition of 
the complete boiler performance, the entire log of readings and data 
should be plotted on a chart and represented graphically. See Fig. 606. 



Lbfc CHART SHOWING LOG OF BOILER TEST. 




M A ,-\\ ' III /' 


/* -~ V eo 


^ .Steam OaugJA * /A |A />' 


I «■>* «/ \ ■X,£v'V' A 


c u tv 2 t3 ^ _ H 'i ^^ 


i»ooo -i \f \*L&3 ^ S 


tx ]£*>*.£ ^ CZ^Sr £ 


' / OEG. V ' 3 


in -3 i£^ i P 




t \- ^\ ^ 


M j H _ | _ ^T? Z> 


^.-c-o -, X- -»"3 ^ r 3 rv 


laoooo s ' .-^V^n /' ^ /\™=J , 


Z ^ \. 2 i/ L I 


• w ■ Z ^1 5 t ^ 5 II S^ 


Feed Temp \ 1 _S* \ / \ 




V £/"' 5 1 j 




±f r ^ f Li*. 


r mi V' ^ 






^^ V^V t"r" 5S,00<> 




/"' \-^'° 




£_ - '■' 20,000 


a 




•^ . 


^ *"' 


c <fra 


S \s 


•S ^^ 


sr ' 


d ^ 


*• m , 


OEG. 2 / 




«•«» gi-ftxtt — ->*'- 


^t~^\ 5£ — J" 12.000 


BO.000 °2:s--:=:.<,.i--^2fp:=: =2 = =2=^9.3 


^_=i==2==^^f — __-5|-^ 10,000 


4>0GO -4 " ^ -'"'' 


draft . . ' Sb " n \ < s-s~"" u ^— -r>— « onm 


7~ "^ -o^ 7 =r '^."' = ^ 




-un, V "^ „£^-., ^~~ 








^ e:"^ 


4000 


? ^^ 


' 


-"i-^' 




^ "' 




. _-7-± - - 





8 9 10 11 12 1 2 3 4 5 C 7 8 9 10 11 12 1 2 3 4 5 C 7 St 



£ 



Fig. 606. 



17. TESTS WITH OIL AND GAS FUELS. 

Tests of boilers using oil or gas for fuel should accord with the rules 
here given, excepting as they are varied to conform to the particular 
characteristics of the fuel. The duration in such cases may be reduced, 
and the " flying" method of starting and stopping employed. 

The table of data and results should contain items stating character of furnace 
and burner, quality and composition of oil or gas, temperature of oil, pressure 
of steam used for vaporizing, and quantity of steam used both for vaporizing 
and for heating. 

TABLE 1. DATA AND RESULTS OF EVAPORATIVE TEST, 
SHORT FORM, CODE OF 1912. 

(1) Test of boiler located at 

to determine conducted by 

(2) Kind of furnace 

(3) Grate surface sq. ft. 

(4) Water-heating surface sq. ft. 

(5) Superheating surface sq. ft. 

(6) Date . 

(7) Duration hr. 

(8) Kind and size of coal 



922 STEAM POWER PLANT ENGINEERING 

Average Pressures, Temperatures, etc. 
(9) Steam pressure by gage lb. 

(10) Temperature of feed water entering boiler deg. 

(11) Temperature of escaping gases leaving boiler deg. 

(12) Force of draft between damper and boiler in. 

(13) Percentage of moisture in steam, or number of degrees of superheat- 

ing per cent or deg. 

Total Quantities. 

(14) Weight of coal as fired* lb. 

(15) Percentage of moisture in coal per cent. 

(16) Total weight of dry coal consumed lb. 

(17) Total ash and refuse lb. 

(18) Percentage of ash and refuse in dry coal per cent. 

(19) Total weight of water fed to the boiler f lb. 

(20) Total water evaporated, corrected for moisture in steam lb. 

(21) Total equivalent evaporation from and at 212 degrees lb. 

Hourly Quantities and Rates. 

(22) Dry coal consumed per hour lb . 

(23) Dry coal per square foot of grate surface per hour lb. 

(24) Water evaporated per hour corrected for quality of steam lb. 

(25) Equivalent evaporation per hour from and at 212 degrees lb. 

(26) Equivalent evaporation per hour from and at 212 degrees per square foot of 

water-heating surface lb. 

Capacity. 

(27) Evaporation per hour from and at 212 degrees (same as Line 25) lb. 

(28) Boiler horse power developed (Item 27 -J- 34|) bl.h.p. 

(29) Rated capacity, in evaporation from and at 212 degrees per hour lb. 

(30) Rated boiler horse power bl.h.p. 

(31) Percentage of rated capacity developed per cent. 

Economy Results. 

(32) Water fed per pound of coal fired (Item 19 -s- Item 14) lb. 

(33) Water evaporated per pound of dry coal (Item 20 -r- Item 16) lb. 

(34) Equivalent evaporation from and at 212 degrees per pound of dry coal 

(Item 21 + Item 16) lb. 

(35) Equivalent evaporation from and at 212 degrees per pound of combustible 

[Item 21 -h (Item 16 - Item 17)] lb. 

Efficiency. 

(36) Calorific value of 1 pound of dry coal B.t.u. 

(37) Calorific value of 1 pound of combustible B.t.u. 

(38) Efficiency of boiler, furnace, and grate 100 X — — -ft q~ — — • • • .per cent. 

(39) Efficiency of boiler and furnace 100 X yr 07 — ~" P er cent. 

* The term " as fired " means actual conditions, including moisture, corrected for estimated differ- 
ence in weight of coal on the grate at beginning and end. 

t Corrected for inequality of water level and steam pressure at beginning and end; 



APPENDIX B 923 

Cost of Evaporation. 

<40) Cost of coal per ton of ... . pounds delivered in boiler room dollars 

(41) Cost of coal required for evaporating 1000 pounds of water from and at 

212 degrees dollars 

TABLE 2. DATA AND RESULTS OF EVAPORATIVE TEST, 
COMPLETE FORM, CODE OF 1912. 

(1) Test of boiler located at 

to determine , conducted by 



Dimensions, Proportions, etc. 

(2) Number and kind of boilers 

(3) Kind of furnace 

(4) Grate surface width length area sq. ft. 

(5) Approximate width of air spaces in grate in. 

(6) Proportion of air space to whole grate surface per cent. 

(7) Water-heating surface sq. ft. 

(8) Superheating surface sq. f t. 

(9) Ratio of water-heating surface to grate surface to 1 

(10) Ratio of minimum draft area to grate surface 1 to 

(11) Date 

(12) Duration hr. 

(13) Kind of coal 

(14) Size of coal 

Average Pressures, Temperatures, etc. 

(15) Steam pressure by gage lb. 

(16) Barometric pressure lb. 

(17) Force of draft at dampers of individual boilers in. 

(18) Force of draft in main flue near boilers in. 

(19) Force of draft in main flue between economizer and chimney in. 

(20) Force of draft in furnaces in. 

(21) Force of blast in ashpits in. 

(22) State of weather 

(23) Temperature of external air deg. 

(24) Temperature of fireroom deg. 

(25) Temperature of steam deg. 

(26) Normal temperature of saturated steam deg. 

(27) Temperature of feed water entering flue heater or economizer deg. 

(28) Temperature of feed water leaving heater or economizer and entering 

boilers deg. 

(29) Temperature of gases lea ving # boilers deg. 

(30) Temperature of gases leaving economizer deg. 

Quality of Steam. 

(31) Percentage of moisture in steam per cent. 

(32) Number of degrees of superheating deg. 

(33) Quality of steam (dry steam = unity) 



924 STEAM POWER PLANT ENGINEERING 

Total Quantities. 

(34) Weight of coal as fired* lb. 

(35) Percentage of moisture in coal per cent. 

(36) Total weight of dry coal consumed lb. 

(37) Total ash and refuse lb. 

(38) Total combustible consumed (Line 36 — Line 37) lb. 

(39) Percentage of ash and refuse in dry coal per cent. 

(40) Total weight of water fed to boiler lb. 

(41) Total water evaporated corrected for moisture in steam lb. 

(42) Factor of evaporation, based on temperature of water entering boilers 

(43) Total equivalent evaporation from and at 212 degrees lb. 

Hourly Quantities and Rates. 

(44) Dry coal consumed per hour lb. 

(45) Combustible consumed per hour lb. 

(46) Dry coal per square foot of grate surface per hour lb. 

(47) Water evaporated per hour, corrected for quality of steam lb. 

(48) Equivalent evaporation per hour from and at 212 degrees! lb. 

(49) Equivalent evaporation per hour from and at 212 degrees per square foot 

of water-heating surface J lb. 

Proximate Analysis of Coal. 

(50) Fixed carbon per cent. 

(51) Volatile matter per cent. 

(52) Moisture per cent. 

(53) Ash per cent. 

100 per cent. 

(54) Sulphur, separately determined per cent. 

Ultimate Analysis of Dry Coal. 

(55) Carbon (C) per cent. 

(56) Hydrogen (H) , per cent. 

(57) Oxygen (O) per cent. 

(58) Nitrogen (N) per cent. 

(59) Sulphur (S) per cent. 

(60) Ash per cent. 

100 per cent. 

(61) Moisture in sample of coal as received per cent. 

Analysis of Ash and Refuse. 

(62) Carbon per cent. 

(63) Earthy matter per cent. 

(a) Si0 2 

(6) A1 2 3 and Fe 2 3 

(c) MgO 

(d) CaO 

(64) Temperature of fusion of ash deg. 

* The term " as fired " means actual conditions, including moisture, corrected for estimated differ- 
ence in weight of coal on the grate at beginning and end. 

t Corrected for inequality of water level and steam pressure at beginning and end. 

t The symbol U. E., meaning Units of Evaporation, may be substituted for the expression, Equiva- 
lent water evaporated into dry steam from and at 212 deg. 



APPENDIX B 925 

Calorific Value. 

(65) Calorific value of 1 pound of dry coal by calorimeter B.t.u. 

(66) Calorific value of 1 pound of combustible by calorimeter B.t.u. 

(67) Calorific value of 1 pound of dry coal by analysis B.t.u. 

(68) Calorific value of 1 pound of combustible by analysis B.t.u. 

Capacity. 

(69) Evaporation per hour from and at 212 degrees (same as Line 48) lb. 

(70) Boiler horse power developed (Line 69 -4- 34|) bl.h.p. 

(71) Rated capacity per hour, from and at 212 degrees % lb. 

(72) Rated boiler horse power bl.h.p. 

(73) Percentage of rated capacity developed per cent. 

Economy Results. 

(74) Water fed per pound of coal (Item 40 -s- Item 34) lb. 

(75) Water evaporated per pound of dry coal (Item 41 -f- Item 36) lb. 

(76) Equivalent evaporation from and at 212 degrees per pound of coal fired 

(Item 43 4- Item 34) lb. 

(77) Equivalent evaporation from and at 212 degrees per pound of dry coal 

(Item 43 -^ Item 36) lb. 

(78) Equivalent evaporation from and at 212 degrees per pound of combustible 

(Item 43 -h Item 38) .lb. 

Efficiency. 

r Item 77 X 970 4T 

(79) Efficiency of boiler, furnace, and grate 100 X ^r w^ — — .... per cent. 

(80) Efficiency of boiler and furnace 100 X ji ™ — — per cent. 

Cost of Evaporation. 

(81) Cost of coal per ton of .... pounds delivered in boiler room .dollars 

(82) Cost of coal required for evaporating 1000 pounds of water under observed 

conditions dollars 

(83) Cost of coal required for evaporating 1000 pounds of water from and at 

212 degrees dollars 

Smoke Data. 

(84) Percentage of smoke as observed per cent. 

(85) Weight of soot per hour obtained from smoke meter 

Methods of Firing. 

(86) Kind of firing, whether spreading, alternate, or cooking 

(87) Average thickness of fire 

(88) Average intervals between firings for each furnace during time when fires 

are in normal condition 

(89) Average interval between times of leveling or breaking up 



926 



STEAM POWER PLANT ENGINEERING 



(90) 
(91) 
(92) 
(93) 
(94) 



Analysis of Dry Gases by Volume. 

Carbon dioxide (C0 2 ) per cent. 

Oxygen (O) per cent. 

Carbon monoxide (CO) per cent. 

Hydrogen and hydrocarbons per cent. 

Nitrogen, by difference (N) per cent. 

100 per cent. 

Heat Balance, Based on Dry Coal and Combustible. 





Dry Coal 
as Fired 


Combustible 
Burned 


B.t.u. 


Per Cent 


B.t.u. 


Per Cent 


(95) Heat absorbed by the boiler (Line 76 or 77 X 970.4) . 










(97) Loss due to heat carried away by steam formed by 


(98) Loss due to heat carried away in the dry flue gases . . 






(102) Loss due to unconsumed hydrogen and hydro-car- 




( 1 03) Total calorific value of 1 lb. of dry coal or combustible 
(Lines 65 and 66) 




100 




100 








APPENDIX C. 

RULES FOR CONDUCTING TESTS OF RECIPROCATING ENGINES.* 
A.S.M.E. Code of 1912. 

1. OBJECT AND PREPARATIONS. 

Determine the object, take the dimensions, note the physical con- 
ditions not only of the engine but of all parts of the plant that are 
concerned in the determinations, examine for leakages, install the 
testing appliances, etc., as pointed out in Appendix A and prepare for 
the test accordingly. 

2. APPARATUS AND INSTRUMENTS. 

The apparatus and instruments required for a simple performance 
test of a steam engine, in which the steam consumption is determined 
by feed-water measurement, are: 

(a) Tanks and platform scales for weighing water (or water meters calibrated 

in place). 

(b) Graduated scales attached to the water glasses of the boilers 

(c) Pressure gauges, vacuum gauges, and thermometers. 

(d) A steam calorimeter. 

(e) A barometer. 

(/) Steam-engine indicators. 

(g) A planimeter. 

(h) A tachometer or other speed-measuring apparatus. 

(i) A friction brake or dynamometer. 

The determination of the heat and steam consumption of an engine 
by feed-water test requires the measurement of the various supplies of 
water fed to the boiler; that of the water discharged by separators and 
drips not returned to the boiler, and that of water and steam which 
escapes by leakage of the boiler and piping, all of these last being 
deducted from the total feed water measured. 

To ascertain the consumption of heat, the various feed temperatures 
are taken and heat calculations made accordingly. If the conditions 
imposed by the particular method adopted for carrying on the test 
depart from the usual practice, as for example where a colder supply of 
feed water is used than the ordinary supply, a preliminary or subsequent 

* Preliminary report of the Committee on Power Tests. (Jour. A.S.M.E., Nov., 
1912.) Greatly abridged. 

927 



928 STEAM POWER PLANT ENGINEERING 

run should be made to ascertain the temperatures which obtain under 
the usual working conditions, and the heat measurements obtained 
under the test conditions appropriately corrected for such departures. 
The steam consumed by steam-driven auxiliaries which are required 
for the operation of the engine should be included in the total steam 
from which the heat consumption is calculated and the quantity of 
steam thus used should be determined and reported. 

3. OPERATING CONDITIONS. 

Determine what the operating conditions should be to conform to 
the object in view, and see that they prevail throughout the trial. 

4. DURATION. 

A test for heat or steam consumption, with substantially constant 
load, should be continued for such time as may be necessary to obtain 
a number of successive hourly records, during which the results are 
reasonably uniform. For a test involving the measurement of feed 
water for this purpose, five hours is sufficient duration. Where a 
surface condenser is used, and the measurement is that of the water 
discharged by the air pump, the duration may be somewhat shorter. 
In this case, successive half-hourly records may be compared and the 
time correspondingly reduced. 

When the load varies widely at different times of the day, the dura- 
tion should be such as to cover the entire period of variation. 

The preliminary or subsequent trial for determining the working tem- 
peratures on a heat test, where the temperatures obtained under the 
test conditions depart from the usual temperatures, should be of such 
duration as may be required to secure working results. 

5. STARTING AND STOPPING. 

The engine and appurtenances having been set to work and thoroughly 
heated under the prescribed conditions of test, except in cases where 
the object is to obtain the performance under working conditions, note 
the water levels in the boilers and feed reservoir, take the time and 
consider this the starting time. Then begin the measurements and 
observations and carry them forward until the end of the period deter- 
mined on. When this time arrives, the water levels and steam pressure 
should be brought as near as practicable to the same points as at the 
start. This being done, again note the time and consider it the stopping 
time of the test. If there are differences in the water levels, proper 
corrections are to be applied. 



APPENDIX C 929 

Where a surface condenser is used, the collection of water discharged 
by the air pump begins at the starting time, and the water is thereafter 
measured or we.ighed until the end of the test; no observations of the 
boilers being required. 

6. RECORDS. 

The general data should be recorded as pointed out in Appendix A, 
under the heading Records. Half-hourly readings of the instruments 
are sufficient, excepting where there are wide fluctuations. A set of 
indicator diagrams should be obtained at intervals of 20 minutes, and 
at more frequent intervals if the nature of the test makes it necessary. 
Mark on each card the cylinder and the end on which it was taken, also 
the time of day. Record on one card of each set the readings of the 
pressure gauges concerned, taken at the same time. These records 
should subsequently be entered on the general log, together with the 
areas, pressures, lengths, etc., measured from the diagrams, when these 
are worked up. 

7. CALCULATION OF RESULTS. 

(a) Dry Steam. The quantity of dry steam consumed when there is no super- 
heating is determined by deducting the moisture found by calorimeter test 
from the total amount of feed water (the latter being corrected for leakages) 
or from the amount of air-pump discharge, as the case may be. 

When there is superheating the dry steam is found by multiplying the 
weight of superheated steam by the factor 

1+ H-h ' 
in which 

C = specific heat of superheated steam at observed pressure and tem- 
perature. 
T = temperature of superheated steam. 
t = temperature of saturated steam. 
H = total heat of saturated steam of observed pressure. 
h = total heat of feed water. 

(6) Heat Consumption. The number of heat units consumed by the engine is 
found by multiplying the weight of feed water consumed, corrected for leak- 
ages, by the total heat of the steam above the working feed temperature, and 
multiplying the product by a factor of correction expressing the quality of 
the steam. 

If the steam contains moisture, this factor equals 

T -h 



Q + P 



H-h 



in which Q is the quality of the steam (one minus the decimal representing 
the percentage of moisture), P the proportion of moisture, T the total heat 



930 STEAM POWER PLANT ENGINEERING 

of water at the temperature of the steam, h the total heat of the feed water, 
and H the total heat of saturated steam. 

If the steam is superheated, the factor is that given above under (a) Dry- 
Steam . 

If there are a number of sources of feed-water supply, the corresponding 
heat units should be determined for each supply and the various quantities 
added together. 

The British standard of heat consumption is based on a feed-water tem- 
perature assumed to be that of the temperature of saturated steam corre- 
sponding to the observed back pressure (whether this is above or below the 
atmosphere), plus the temperature due to heat derived from jacket or reheater 
drips. It does not include the heat consumed by any auxiliaries, except 
jackets and reheaters. 

(c) Indicated Horse Power. In a single double-acting cylinder the indicated 
horse power is found by using the formula 

PLAN 

33,000 ' 

in which P represents the average mean effective pressure in pounds per 
square inch measured from the indicator diagrams, L the length of stroke in 
feet, A the area of the piston less one-half the area of the piston rod, or the 
mean area of the rod if it passes through both cylinder heads, in square inches, 
and N the number of single strokes per minute. 

Where extreme accuracy is required, the power developed by each side of 
the piston may be determined and the results added together. 

(d) Brake Horse Power. The brake horse power is found by multiplying the net 
weight on the brake arm (the gross weight minus the weight when the brake 
is entirely free) in pounds, the circumference of the circle passing through 
the bearing point at the end of the brake arm, in feet, and the number of 
revolutions of the brake shaft per minute, and dividing the product by 33,000. 

(e) Electrical Horse Power. The electrical horse power for a direct-connected 
generator is found by dividing the output at the bus-bar, expressed in kilo- 
watts, by the decimal 0.746. For alternating-current systems the net output 
is to be used, being the total output less that consumed for excitation. 

(J) Efficiency. The efficiency is expressed by the thermal efficiency ratio, 

which is found by dividing the quantity 2545 by the number of heat units 

consumed per h.p. hr., either indicated or brake. 
(g) Steam accounted for by Indicator Diagrams. The steam accounted for, 

expressed in pounds per i.h.p. per hour, may readily be found by using the 

formula 

i^52 [~(C + E) Wc - (H + E) Wh], 
m.e.p. L J 

in which 

m.e.p. = mean effective pressure. 

C = proportion of stroke completed at cut-off or release. 

E = proportion of clearance. 

H = proportion of stroke uncompleted at compression. 

Wc = weight of 1 cubic foot steam at cut-off or release pressure. 

Wh = weight of 1 cubic foot steam at compression pressure. 



APPENDIX C 931 

The points of cut-off release and compression, referred to, are indicated in 
Fig. 607.* 

In multiple-expansion engines the mean effective pressure to be used in 
the above formula is the combined m.e.p. referred to the cylinder under con- 
sideration. In a compound engine the combined m.e.p. for the h.p. cylinder 
is the sum of the actual m.e.p. of the h.p. cylinder and that of the l.p. cy Under 
multiplied by the cy Under ratio. Likewise the combined m.e.p. for the l.p. 
cylinder is the sum of the actual m.e.p. of the l.p. cyUnder and the m.e.p. 
of the h.p. cyUnder divided by the cylinder ratio. 




Compression 



Atmospheric Line 

Fig. 607. Points where "Steam accounted for by Indicator" is Computed. 

(h) Cut-Off and Ratio of Expansion. To find the percentage of cut-off, or 
what may best be termed the "commercial cut-off," the following rule should 
be observed: 

Through the point of maximum pressure during admission draw a 
line paraUel to the atmospheric line. Through a point on the expan- 
sion line where the cut-off is complete, draw a hyperbolic curve. The 
intersection of these two lines is the point of commercial cut-off, and 
the proportion of cut-off is found by dividing the length measured on 
the diagram up to this point by the total length. 

To find the ratio of expansion divide the volume corresponding to the 
piston displacement, including clearance, by the volume of the steam at the 
commercial cut-off, including clearance. 

In a multiple-expansion engine the ratio of expansion is found by dividing 
the volume of the l.p. cylinder, including clearance, by the volume of the 
h.p. cylinder at the commercial cut-off, including clearance. 

8. DATA AND RESULTS. 

The data and results should be reported in accordance with either 
the Short Form or Complete Form given herewith, adding lines for 
data not provided for, or omitting those not required, as may conform 
to the object in view. 

* Reproduced from Report of Committee on Standardizing Engine Tests, Fig. 122, Trans. Am. Soc. 
M. E., vol. 24, p. 744. 



932 STEAM POWER PLANT ENGINEERING 

TABLE 1. DATA AND RESULTS OF HEAT AND FEED- WATER TESTS 

OF STEAM ENGINE. 

SHORT FORM, CODE OF 1912. 

(1) Test of engine located at 

to determine conducted by 

(2) Type and class of engine and auxiliaries 

(3) Dimensions of main engine: 1st Cyl. 2dCyl. 3dCyl. 

(a) Diameter of cylinder in 

(b) Stroke of piston ft ' 

(c) Diameter of piston rod each end in 

(d) Average clearance per cent 

(e) Cylinder ratio 

(/) Horse-power constant for 1 lb. m.e.p. and 

1 r.p.m 

(4) Dimensions and type of auxiliaries 

(5) -Date 

(6) Duration hr. 

Average Pressures and Temperatures. 

(7) Pressure in steam pipe near throttle by gauge lb. 

(8) Barometric pressure of atmosphere in in. of mercury in. 

(9) Pressure in receivers by gauge lb. 

(10) Vacuum in condenser in inches of mercury , in. 

(11) Pressure in jackets and reheaters by gauge . . . . : lb. 

(12) Temperature of main supply of feed water deg. 

(13) Temperature of additional supplies of feed water deg. 

Total Quantities. 

(14) Total water fed to boilers from main source of supply lb. 

(15) Total water fed from additional supplies lb. 

(16) Total water fed to boilers from all sources lb. 

(17) Moisture in steam or superheating near throttle per cent or deg. 

(18) Factor of correction for quality of steam 

(19) Total dry steam consumed for all purposes lb. 

Hourly Quantities. 

(20) Water fed from main source of supply lb. 

(21) Water fed from additional supplies lb. 

(22) Total water fed to boilers per hour lb. 

(23) Total dry steam consumed per hour lb. 

(24) Loss of steam and water per hour due to drips from main steam pipes and 

to leakage of plant lb. 

(25) Net dry steam consumed per hour by engine^and auxiliaries lb. 

(26) Net dry steam consumed per hour: 

(a) By engine alone lb. 

(6) By auxiliaries lb. ; 



APPENDIX C 933 

Heat Data. 

(27) Heat units per pound of dry steam, based on temperature of Line 12 B.t.u. 

(28) Heat units per pound of dry steam, based on temperature of Line 13 B.t.u. 

(29) Heat units consumed per hour, main supply of feed B.t.u. 

(30) Heat units consumed per hour, additional supplies of feed B.t.u. 

(31) Total heat units consumed per hour for all purposes B.t.u. 

(32) Loss of heat per hour due to leakage of plant, drips, etc B.t.u. 

(33) Net heat imits consumed per hour: 

(a) By engine and auxiliaries B.t.u. 

(6) By engine alone B.t.u. 

(c) By auxiliaries B.t.u. 

Indicator Diagrams. 

IstCyl. 2dCyl. 3dCyl. 

(34) Commercial cut-off in per cent of stroke 

(35) Initial pressure in lb. per sq. in. above atmosphere 

(36) Back pressure at lowest point above or below atmos- 

phere in lb. per sq. in • 

(37) Mean effective pressure in lb. per sq. in 

(38) Steam accounted for by indicator in lb. per i.h.p. 

per hour : 

(a) Near cut-off 

(6) Near release 

Speed. 

(39) Revolutions per minute . rev. 

(40) Piston speed in feet per minute ft. 

Power. 

(41) Indicated horse power developed by main-engine cylinders: 

1st cylinder i.h.p. 

2d cylinder i.h.p. 

3d cylinder whole engine .• i.h.p. 

Whole engine i.h.p. 

(42) Brake horse power br.h.p. 

Economy Results. 

(43) Heat units consumed by engine and auxiliaries per hour: 

(a) Per indicated horse power B.t.u. 

(b) Per brake horse power B.t.u. 

(44) Dry steam consumed per indicated horse power per hour: 

(a) By engine and auxiliaries lb. 

(b) By main engine alone lb. 

(c) By auxiliaries lb. 

(45) Dry steam consumed per brake horse power per hour: 

(a) By engine and auxiliaries lb. 

(6) By main engine alone lb. 

(c) By auxiliaries lb. 

(46) Percentage of steam used by main-engine cylinders accounted for by indi- 

cator diagrams: 

(a) Near cut-off per cent. 

(6) Near release per cent. 



934 STEAM POWER PLANT ENGINEERING 



Sample Diagrams. 

TABLE 2. DATA AND RESULTS OF STEAM-ENGINE TEST, COMPLETE 

FORM, CODE OF 1912. 

(1) Test of engine located at 

to determine conducted by 

(2) Type of engine (simple, compound, or other multiple expansion; condensing 

or non-condensing) 

(3) Class of engine (mill, marine, electric, etc.) 

(4) Rated power of engine 

(5) Name of builders 

(6) Number and arrangement of cylinders of engine; how lagged; type of 

condenser 

(7) Type of valves 

(8) Type of boiler 

(9) Kind and type of auxiliaries (air pump, circulating pump, feed pump; 

jackets, heaters, etc.) 

(10) Dimensions of engine: IstCyl. 2dCyl. 3d Cyl. 

(a) Single or double acting 

(&) Cylinder dimensions: 

Bore, in 

Stroke, ft 

Diameter of piston rod, in 

Diameter of tail rod, in 

(c) Clearance in per cent of volume displaced by 

piston per stroke: 

Head end, per cent 

Crank end, per cent 

Average, per cent 

(d) Surface in square feet (average) : 

Barrel of cylinder, square feet 

Cylinder heads, square feet 

Clearance and ports, square feet 

Ends of piston, square feet 

(c) Jacket surfaces or internal surfaces of cylinder 
heated by jackets, in square feet: 

Barrel of cylinder, square feet 

Cylinder heads, square feet 

Clearance and ports, square feet 

Receiver jackets, square feet 

(g) Horse power constant for 1 lb. m.e.p. and 

1 r.p.m 

(11) Dimensions of boilers: 

(a) Number 

(b) Total grate surface sq. ft. 

(c) Total water-heating surface sq. ft. 

(d) Total steam-heating surface sq. ft. 



APPENDIX G 935 

(12) Dimensions of auxiliaries: 

(a) Air pump 

(6) Circulating pump 

(c) Feed pumps 

(d) Heaters 

(13) Dimensions of condenser 

(14) Dimensions of electric or other machinery driven by engine 

(15) Date 

(16) Duration hr. 

Average Pressures and Temperatures. 

(17) Steam pressure at boiler by gauge lb. 

(18) Steam-pipe pressure near throttle, by gauge lb. 

(19) Barometric pressure of atmosphere in lb. per sq. in. . lb. 

(20) Pressure in first receiver by gauge lb. 

(21) Pressure in second receiver by gauge lb. 

(22) Vacuum in condenser: 

(a) In in. of mercury in. 

(6) Corresponding total pressure lb. 

(23) Pressure in steam jacket by gauge lb. 

(24) Pressure in reheater by gauge lb. 

(25) Superheat in steam leaving first receiver deg. 

(26) Superheat in steam leaving second receiver deg. 

(27) Temperature of main supply of feed water to boilers deg. 

(28) Temperature of additional supplies of feed water deg. 

(29) Ideal feed- water temperature corresponding to the pressure of the steam in 

the exhaust pipe, allowance being made for heat derived from jacket or 
reheater drips (British Standard) deg. 

(30) Temperature of injection or circulating water entering condenser deg. 

(31) Temperature of injection or circulating water leaving condenser deg. 

(32) Temperature of air in engine room deg. 

Total Quantities. 

(33) Water fed to boilers from main source of supply lb. 

(34) Water fed from additional supplies lb. 

(35) Total water fed to boilers from all sources lb. 

(36) Moisture in steam or superheating near throttle per cent or deg. 

(37) Factor of correction for quality of steam, dry steam being unity lb. 

(38) Total dry steam consumed for all purposes 

Hourly Quantities. 

(39) Water fed from main source of supply lb. 

(40) Water fed from additional supplies '. lb. 

(41) Total water fed to boilers per hour lb. 

(42) Total dry steam consumed per hour lb. 

(43) Loss of steam and water per hour due to drips from main steam pipes and to 

leakage of plant lb. 

(44) Net dry steam consumed per hour by engine and auxiliaries lb. 



936 STEAM POWER PLANT ENGINEERING 

(45) Dry steam consumed per hour: 

(a) Main engine alone lb. 

(6) Jackets and reheaters lb. 

(c) Air pump lb. 

(d) Circulating pump lb. 

(e) Feed- water pump „. lb. 

(/) Other auxiliaries lb. 

(46) Injection or circulating water supplied condenser per hour cu. ft. 

Heat Data. 

(47) Heat units per pound of dry steam, based on temperature of Line 27 B.t.u. 

(48) Heat units per pound of dry steam, based on temperature of Line 28 B.t.u. 

(49) Heat units consumed per hour, main supply of feed. B.t.u. 

(50) Heat units consumed per hour, additional supplies of feed B.t.u. 

(51) Total heat units consumed per hour for all purposes B.t.u. 

(52) Loss of heat per hour due to leakage of plant, drips, etc. B.t.u. 

(53) Net heat units consumed per hour: 

(a) By engine and auxiliaries B.t.u. 

(b) By engine alone B.t.u. 

(c) By auxiliaries B.t.u. 

(54) Heat units consumed per hour by the engine alone, reckoned from tem- 

perature given in Line 29 (British Standard) B.t.u. 

Indicator Diagrams. 

IstCyl. 2dCyl. 3d Cyl. 

(55) Commercial cut-off in per cent of stroke 

(56) Initial pressure in lb. per sq. in. above atmosphere 

(57) Back-pressure at mid-stroke above or below atmos- 

phere in lb. per sq. in 

(58) Mean effective pressure in lb. per sq. in 

(59) Equivalent mean effective pressure in lb. per sq. in. : 

(a) Referred to first cylinder 

(b) Referred to second cylinder 

(c) Referred to third cylinder 

(60) Pressures and percentages used in computing the 

steam accounted for by the indicator diagrams, 
measured to points on the expansion and com- 
pression curves: 

Pressure above zero in lb . per sq. in. : 

(a) Near cut-off 

(6) Near release 

(c) Near beginning of compression 

Percentage of stroke at points where pressures 

are measured: 

(d) Near cut-off 

(e) Near release 

(/) Near beginning of compression 

(61) Aggregate m.e.p. in lb. per sq. in. referred to each 

cylinder given in heading 

(62) Mean back pressure above zero, lb. per sq. in. . 



APPENDIX C 937 

(63) Steam accounted for in lb. per indicated horse power 1st Cyl. 2d Cyl. 3d Cyl. 

per hour: 

(a) Near cut-off 

(6) Near release 

(64) Ratio of expansion 

(65) Mean effective pressure of ideal diagram lb. 

(66) Diagram factor 

Speed. 

(67) Revolutions per minute rev. 

(68) Piston speed per minute ft. 

(69) Variation of speed between no load and full load rev. 

(70) Fluctuation of speed on suddenly changing from full load to no load, measured 

by the increase in the revolutions due to the change rev. 

Power. 

(71) Indicated horse power developed by main engine: 

1st cylinder i.h.p. 

2d cylinder i.h.p. 

3d cylinder ... i.h.p. 

Whole engine i.h.p. 

(72) Brake horse power br.h.p. 

(73) Friction i.h.p. by diagrams, no load on engine, computed for average speed . i.h.p. 

(74) Difference between Lines 71 and 72 h.p. 

(75) Percentage of i.h.p. of main engine lost in friction per cent. 

(76) Power developed by auxiliaries* i.h.p. 

Economy Results. 

(77) Heat units consumed per indicated horse power per hour:f 

(a) By engine and auxiliaries B.t.u. 

(6) By engine alone B.t.u. 

(78) Heat units consumed per brake horse power per hour: 

(a) By engine and auxiliaries B.t.u. 

(6) By engine alone B.t.u. 

(79) Heat units consumed by engine per hour, corresponding to ideal tempera- 

ture of feed water given in Line 29, per indicated horse power (British 
Standard) B.t.u. 

(80) Dry steam consumed per i.h.p. per hour: 

(a) By engine and auxiliaries lb. 

(6) By main engine alone lb. 

(c) By auxiliaries lb. 

* These are not included in the power developed by the main engine. 

t The h.p. on which the economy and efficiency results are based are those of the main engine 
given in Line 71. 

Note: Both the Short Form and Complete Form here given refer to a steam engine used for gen- 
eral service. 

For an engine driving an electric generator the form should be enlarged to include the electrical 
data, embracing the average voltage, number of amperes each phase, number of watts, number of 
watt-hours, average power factor, etc.; and the economy results based on the electrical output em- 
bracing the heat units and steam consumed per electric h.p. per hour and per kw.-hr., together with 
the efficiency of the generator. See table for Steam Turbine Code, Part 6. 

Likewise, in a marine engine having a shaft dynamometer, the form should include the data ob- 
tained from this instrument, in which case the brake h.p. becomes the shaft h.p. 



938 STEAM POWER PLANT ENGINEERING 

(81) Dry steam consumed per brake h.p. per hour: 

(a) By engine and auxiliaries lb. 

(b) By main engine alone lb. 

(c) By auxiliaries lb. 

(82) Percentage of steam used by main engine cylinders accounted for by indicator 

diagrams. 

IstCyl. 2dCyl. 3d Cyl. 

(a) Near cut-off 

(6) Near release 

Efficiency Results. 

(83) Thermal efficiency ratio for engine and auxiliaries: 

(a) Per indicated horse power per cent. 

(b) Per brake horse power per cent. 

(84) Thermal efficiency ratio for engine alone: 

(a) Per indicated horse power per cent. 

(6) Per brake horse power per cent. 

(85) Ratio of economy of engine to that of an ideal engine working with the 

Rankine cycle per cent. 

Work Done Per Heat Unit. 

(86) Ft. -lb. of net work per B.t.u. consumed by engine and auxiliaries (1,980,000 

-7- Line 78a) ft.-lb. 






APPENDIX D. 

RULES FOR CONDUCTING TESTS OF STEAM TURBINES AND 
TURBO-GENERATORS.* 

A.S.M.E. Code of 1012. 

1. OBJECT AND PREPARATIONS. 

Determine the object, take the dimensions, note the physical con- 
ditions not only of the turbine but of the entire plant concerned, ex- 
amine for leakages, install the testing appliances, etc., as pointed out 
in the general instructions in Appendix A, and prepare for the test 
accordingly. 

2. APPARATUS AND INSTRUMENTS. 

The apparatus and instruments required for a simple performance test 
of a steam turbine or turbo-generator, in which the steam consumption 
is determined by feed-water measurement, are : 

(a) Tanks and platform scales for weighing water (or water meters calibrated in 
place). 

(6) Graduated scales attached to the water glasses of the boilers. 

(c) Pressure gauges, vacuum gauges, and thermometers. 

(d) A steam calorimeter. 

(e) A barometer. 

(/) A tachometer or other speed-measuring apparatus. 
(g) A friction brake or dynamometer. 

(h) Voltmeters, ammeters, wattmeters, and watt-hour meters for the electrical 
measurements in the case of a turbo-generator. 

The determination of the heat and steam consumption of a turbine 
or turbo-generator should conform to the same methods as those de- 
scribed in the Steam Engine Code. 

The steam consumed by steam-driven auxiliaries required for the 
operation of a turbine should be included in the total steam from which 
the heat consumption is calculated the same as in the case of the steam 
engine. 

3. OPERATING CONDITIONS. 

Determine what the operating conditions should be to conform to 
the object in view and see that they prevail throughout the trial. 

* Preliminary Report of Committee on Power Tests. (Jour. A.S.M.E., Nov., 1912.) 
Somewhat abridged. 



940 STEAM POWER PLANT ENGINEERING 

4. DURATION. 
5. STARTING AND STOPPING. 

6. RECORDS. 
7. CALCULATION OF RESULTS. 

The rules pertaining to the subjects Duration, Starting and Stopping, 
Records, and Calculation of Results are identically the same as those 
given under the respective headings in the Steam Engine Code, with 
the single exception of the matter relating to indicator diagrams and 
results computed therefrom; and reference may be made to that code 
for the directions required in these particulars. 

8. DATA AND RESULTS. 

The data and results should be reported in accordance with the form 
given herewith, addhig lines for data not provided for, or omitting those 
not required, as may conform to the object in view: 

DATA AND RESULTS OF STEAM TURBINE OR TURBO-GENERATOR 

TESTS. 



Average Pressures and Temperatures. 

(13) Steam-pipe pressure near throttle, by gauge lb. 

(14) Steam-chest pressure by gauge lb. 

(15) Barometric pressure of atmosphere in lb. per sq. in lb. 

(16) Vacuum in condenser: 

(a) In inches of mercury in. 

(b) Corresponding absolute pressure lb. 

(17) Exhaust-chamber pressure (absolute) lb. 

(18) Temperature of main supply of feed water to boilers deg. 

(19) Temperature of additional supplies of feed water deg. 

(20) Temperature of injection or circulating water entering condenser deg. 

(2-1) Temperature of injection or circulating water leaving condenser deg. 






CODE OF 1912. 

(1) Test of turbine located at 

to determine , conducted by 

(2) Type of turbine and class of service 

(3) Type of generator, kind of current, etc ' 

(4) Rated power of turbine 

(5) Type of boiler 

(6) Kind and type of auxiliaries (air pumps, circulating pumps, feed pumps, etc.) 

(7) Dimensions of turbine or turbo-generator 

(8) Dimensions of boilers 

(9) Dimensions of auxiliaries 

(10) Dimensions of condenser 

(11) Date 

(12) Duration hr. 



I 
I 



APPENDIX D 941 

Total Quantities. 

(22) Water fed to boilers from main source of supply lb. 

(23) Water fed from additional supplies lb. 

(24) Total water fed to boilers from all sources lb. 

(25) Moisture in steam or superheating near throttle per cent or deg. 

(26) Factor of correction for quality of steam, dry steam being unity 

(27) Total dry steam consumed for all purposes lb. 

Hourly Quantities. 

(28) Water fed from main source of supply lb. 

(29) Water fed from additional supplies lb. 

(30) Total water fed to boilers per hour lb. 

(31) Total dry steam consumed per hour lb. 

(32) Loss of steam and water per hour due to drips from main steam pipes and to 

leakage of plant lb. 

(33) Net dry steam consumed per hour lb. 

(34) Dry steam consumed per hour: 

(a) By turbine lb. 

(b) By auxiliaries lb. 

(35) Injection or circulating water supplied condensers per hour cu. ft. 

Heat Data. 

(36) Heat units per pound of dry steam, based on temperature of Line 18 B.t.u. 

(37) Heat units per pound of dry steam, based on temperature of Line 19 B.t.u. 

(38) Heat units consumed per hour, main supply of feed B.t.u. 

(39) Heat units consumed per hour, additional supplies of feed .B.t.u. 

(40) Total heat units consumed per hour for all purposes B.t.u. 

(41) Loss of heat per hour due to leakage of plant, drips, etc .B.t.u. 

(42) Heat units consumed per hour: 

(a) By turbine and auxiliaries B.t.u. 

(6) By turbine alone B.t.u. 

(c) By auxiliaries B.t.u. 

Electrical Data. 

(43) Average volts, each phase volts. 

(44) Average amperes, each phase amperes. 

(45) Average kilowatts, first meter kw. 

(46) Average kilowatts, second meter kw. 

(47) Total kilowatt output kw. 

(48) Power factor 

(49) Output consumed by exciter kw. 

(50) Net kilowatt output kw. 

Speed. 

(51) Revolutions per minute rev. 

(52) Variation of speed between no load and full load rev. 

(53) Fluctuation of speed on suddenly changing from full load to no load, 

measured by the increase in the revolutions due to the change rev. 

Power. 

(54) Brake horse power br.h.p. 

(55) Electrical horse power , h.p. 






942 STEAM POWER PLANT ENGINEERING 

Economy Results. 

(56) Heat units consumed by turbine and auxiliaries per brake h.-p. hr B.t.u. 

(57) Dry steam consumed per brake h.-p. hr. : 

(a) By turbine and auxiliaries lb. 

(b) By turbine alone , lb. 

(c) By auxiliaries lb. 

(58) Dry steam consumed per kw.-hr.: 

(a) By turbine and auxiliaries lb. 

(6) By turbine alone lb. 

(c) By auxiliaries lb. 

Efficiency Results. 

(59) Thermal efficiency ratio per brake horse power . . . per cent. 

(60) Ratio of economy of turbine to that of an ideal turbine working with the 

Rankine cycle 

Work Done per Heat Unit. 

(61) Ft. -lb. of net work per B.t.u. consumed by turbine and auxiliaries (1,980,000 

4- Line 56) ft.-lb. 



APPENDIX E. 

RULES FOR CONDUCTING TESTS OF COMPLETE STEAM POWER 

PLANTS.* 

A.S.M.E. Code of 1912. 

1. OBJECT AND PREPARATIONS. 

These rules are intended to apply to commercial tests of a complete 
plant to determine the number of pounds of fuel consumed per unit of 
work done in a unit of time. For tests of the component parts of a 
complete plant, such as boilers, engines, turbines, etc., rules may be 
found in there spective Codes applying to such cases. 

Read the general instructions given in Appendix A. Take the 
dimensions, note the physical conditions, examine for leakages, install 
the testing appliances, etc., as there pointed out, and prepare for the 
test accordingly. 

2. FUEL. 

Determine the character of the fuel to be used according to the object 
In view. For further particulars reference may be made to the Boiler 
Code. 

3. APPARATUS AND INSTRUMENTS. 

The apparatus and instruments required for a simple performance 
test of a steam plant are: 

(a) Platform scales for weighing coal and ashes. 
(6) Coal calorimeter. 

(c) Steam-engine indicators. 

(d) A tachometer or other speed-measuring apparatus. 

(e) Electrical instruments for determining the output of an electric plant. 

If the test involves the determination of boiler performance, and 
engine or turbine performances, additional instruments should be used 
as pointed out in the respective Codes referring to such tests. 

4. OPERATING CONDITIONS. 

Determine what the operating conditions should be to conform to 
the object in view, and see that they prevail throughout the trial. 

* Preliminary Report of the Committee on Power Tests. (Jour. A.S.M.E., Nov., 
1912.) 

943 



944 STEAM POWER PLANT ENGINEERING 

5. DURATION. 

The duration of a plant test should be not less than one day of 24 
hours, and preferably a full week of seven days, including Sunday. 

In cases where the engine or turbine is in operation only a part of 
the day, the duration on which the results are computed, should be 
considered the length of time that the engine or turbine is in opera- 
tion at its working speed. 

6. STARTING AND STOPPING. 

In a plant operating continuously, day and night, the times fixed for 
starting and stopping should follow the regular periods of cleaning 
the fires. The fires should be quickly cleaned and then burned low, 
say to a thickness of 4 inches. When this condition is reached the 
time should be noted as the starting time, and the thickness of each 
coal bed observed, as also the water levels and the steam pressure. 
Fresh coal should then be fired from that weighed for the test, the 
ashpits thoroughly cleaned, and the regular work of the test proceeded 
with. At the close of the test, following a regular cleaning, the fires 
should again be burned low, and when their condition has become the 
same as that observed at the beginning, the water levels and steam 
pressure also being the same, the time is observed and this time taken 
as the stopping time. If the water levels and steam pressure are not 
the same as at the beginning a suitable correction should be made by 
computation. The ashes and refuse are then hauled from the ashpits. 

In a plant running only a part of the day, and during the balance of 
the day the fires are banked, the time selected for the beginning and 
end of the test should be that following the close of the day's run, when 
the fires have been burned low preparatory to cleaning and banking. 
The amount of live coal left on the grates under these circumstances is 
estimated at the beginning of the test, and the fires brought to the same 
condition, as near as may be, at the close of the test the next day. If 
the two quantities differ, a suitable correction is made in the weight 
of coal fired, as found by calculation. 

7. RECORDS. 
The general data should be recorded as pointed out in Appendix A, 
under the head of Records. Half-hourly readings of the various 
instruments concerned are usually sufficient, excepting where there 
are wide fluctuations. A set of indicator diagrams should be obtained 
at intervals of 20 minutes, and at more frequent intervals if the nature 
of the test makes it necessary. Mark on each card the cylinder and 
the end on which it was taken, also the time of the day. Record on 



APPENDIX E 945 

one card of each set the readings of the pressure gauges concerned, taken 
at the same time. These records should subsequently be entered on 
the general log, together with the areas, pressures, lengths, etc., meas- 
ured from the diagrams, when these are the worked up. 

8. SAMPLING AND DRYING COAL. 
During the progress of the test the coal should be regularly sampled 
for the purpose of analysis and determination of moisture. 

9. ASHES AND REFUSE. 
The ashes and refuse withdrawn from the furnace and ashpit during 
the progress of the test and at its close should be weighed in a dry 
state, and, if desired, a representative sample should be obtained for 
proximate analysis and the determination of the amount of unburned 
carbon which it contains. 

10. CALORIFIC TESTS AND ANALYSES OF COAL. 
The quality of the fuel should be determined by calorific tests and 
analysis of the representative sample above referred to. 

11. CALCULATION OF RESULTS. 
The methods of calculating the indicated and electrical horse powers 
are the same as explained in the Steam Engine Code. 

12. DATA AND RESULTS. 
The data and results should be reported in accordance with the form 
given herewith, adding lines for data not provided for, or omitting those 
not required, as may conform to the object in view. 

DATA AND RESULTS OF COMPLETE STEAM POWER PLANT TEST. 

(1) Test of plant located at 

to determine , conducted by 

(2) Type of engine or turbine and class of service 

(3) Rated power of engine or turbine 

(4) Type of boilers 

(5) Kind and type of auxiliaries (air pump, circulating pump, and feed pump; 

jackets, heaters, etc.) 

(6) Dimensions of engine or turbine 

(7) Dimensions of boilers ; 

(8) Dimensions of auxiliaries 

(9) Dimensions of condenser 

(10) Date 

(11) Duration .hr. 

(12) Length of time engine or turbine was in motion with throttle open hr. 

(13) Length of time engine or turbine was running at normal speed hr. 

(14) Kind of coal 

(15) Size of coal 



946 STEAM POWER PLANT ENGINEERING 

Average Pressures and Temperatures. 

(16) Steam pressure at boiler by gauge .' lb. 

(17) Steam-pipe pressure near throttle, by gauge lb. 

(18) Barometric pressure of atmosphere in in. of mercury in. 

(19) Pressure in receiver by gauge lb. 

(20) Vacuum in condenser in. 

(21) Number of degrees of superheating, if any, near throttle deg. 

(22) Temperature of feed water entering boilers deg. 

Total Quantities, Time, Etc. 

(23) Total coal as fired * lb. 

(24) Moisture in coal per cent. 

(25) Total dry coal consumed lb. 

(26) Ash and refuse lb. 

(27) Percentage of ash and refuse to dry coal per cent. 

(28) Calorific value by calorimeter test per lb. of dry coal B.t.u. 

(29) Cost of coal per ton of .... lb dollars. 

Hourly Quantities. 

(30) Dry coal consumed per hour, based on duration of running period lb. 

Indicator Diagrams. 

(31) Mean effective pressure in lb. per sq. in lb. 

Electrical Data. 

(32) Total electrical output kw.-hr. 

(33) Electrical output per hour kw. 

(34) Output consumed by exciter kw. 

(35) Net electrical output per hour kw. 

(36) Average volts each phase volts. 

(37) Average amperes each phase amperes. 

(38) Power factor 

Speed. 

(39) Revolutions per minute rev. 






Power. 

(40) Indicated horse power developed by main engine: 

First cylinder i.h.p. 

Second cylinder i.h.p. 

Whole engine i.h.p. 

(41) Net electrical horse power h.p. 

Economy Results. 

(42) Dry coal consumed per i.h.p. per hour lb. 

(43) Dry coal consumed per kw.-hr lb. 

(44) Cost of coal per i.h.p. per hour cents. 

(45) Cost of coal per kw.-hr cents. 

* When an independent superheater is used, this includes coal burned in the superheater. 



APPENDIX F. 

RULES FOR CONDUCTING DUTY TRIALS OF STEAM PUMPING 

MACHINERY.* 

A.S.M.E. Code of 1912. 

1. OBJECT AND PREPARATIONS. 

Read the general instructions given in Appendix A. Determine the 
object, take the dimensions, note the physical conditions not only of 
the pumping machinery but of all parts of the plant concerned, examine 
for leakages, install the testing appliances, etc., as there pointed out, 
and prepare for the test accordingly. 

In a reciprocating pump, determine the quantity of water leakage 
or slip past the plungers, and that of the pump valves, if any. 

2. APPARATUS AND INSTRUMENTS. 

The apparatus and instruments required for a simple duty trial of 
pumping machinery, in which the steam consumption is determined 
by feed- water measurement, are : 

(a) Tanks and platform scales for weighing water (or water meters calibrated in 

place). 
(6) Graduated scales attached to the water glasses of the boilers. 

(c) Pressure gauges, vacuum gauges, and thermometers. 

(d) A steam calorimeter. 

(e) A barometer. 

(/) A tachometer or other speed-measuring apparatus. 

(g) For rotary pumps a weir or other means for measuring the quantity of water 

pumped. 
(h) Stroke scales, for direct-acting pumps. 

In trials of a reciprocating pumping engine, involving the determina- 
tion of the complete performance, the weir or other means of measure- 
ment noted should be provided, and in addition the following: 

00 Steam-engine indicators, 
(j) A planimeter. 

* Preliminary report of the Committee on Power Tests. (Jour. A.S.M.E., Nov., 
1912.) In the case of a pump driven by some other prime mover than a steam engine 
or steam turbine, the code may be modified to suit the particular circumstances. 

947 



948 STEAM POWER PLANT ENGINEERING 

The determination of the heat and steam consumption should conform 
to the same methods as those described in the Steam Engine Code. 

The steam consumed by steam-driven auxiliaries which are required 
in the operation of the pumping machinery should be included in the 
total steam from which the heat consumption is calculated, the same 
as noted in the Steam Engine Code. 

3. OPERATING CONDITIONS. 

Determine what the operating conditions should be to conform to 
the object in view and see that they prevail throughout the trial. 

In trials for maximum duty, care should be taken that no air is 
snifted into the pump cylinders, causing imperfect filling. In such 
cases, and indeed in all cases where air is thus admitted in sufficient 
quantity to affect the performance as revealed by indicator diagrams 
from the water end, the result should be corrected accordingly. 

4. DURATION. 

5. STARTING AND STOPPING. 

6. RECORDS. 

The rules pertaining to the subjects Duration, Starting and Stopping 
and Records are identically the same as those given under the respective 
headings in the Steam Engine Code, and reference may be made to 
that code for the necessary directions in these particulars. Where the 
pump end is of the reciprocating class, the indicator diagrams should 
be taken not only from the steam cylinders but also from the water 
cylinders. 

7. CALCULATION OF RESULTS. 

The rules pertaining to Dry Steam, Heat Consumption, and In- 
dicated Horse Power are identically the same as those given in the 
Steam Engine Code; and reference may be made to that code for the 
necessary directions in these particulars. 

(a) Water Horse Power. The water horse power in a reciprocating pump is 
found by multiplying the net area of the plunger in square inches by the total 
head, which is made up of the pressure shown by the gauge on the force main, 
that on the suction main, and that representing the vertical distance between 
the centers of the two gauges, all expressed in pounds per square inch; the 
length of the stroke in feet; and the number of single strokes per minute; 
and dividing the final product by 33,000. 

In a rotary pump the water horse power is found by multiplying the weight 
of water discharged per hour in pounds, as determined by weir or other 



APPENDIX F 949 

measurement; by the total head in feet, as determined from the readings of the 
gauge on the force main, that on the suction main, and the vertical distance 
between the two gauges; and the product divided by 1,980,000.* 

(b) Duty. The duty per million heat units is found by dividing the number 
of foot-pounds of work done during the trial by the total number of heat units 
consumed; and multiplying the quotient by 1,000,000. The amount of work 
is found in the case of reciprocating pumps by multiplying the net area of the 
plunger in square inches, the total head expressed in pounds per square inches 
(which is made up of the pressure shown by the gauge on the force main, that 
on the suction main, and the vertical distance between the centers of the two 
gauges, all reduced to pounds), by the length of the stroke in feet, and the 
total number of single strokes during the trial; finally correcting for the per- 
centage of leakage of the pump. In a rotary pump the work done is found by 
multiplying the weight of water discharged during the trial, as determined by 
weir or other measurement, by the total head in feet. 

The duty per 1000 pounds of dry steam is found by dividing the foot- 
pounds of work done, as noted above, by the total weight of dry steam, and 
multiplying the quotient by 1000. 

(c) Capacity. The capacity in gallons per 24 hours for reciprocating pumps is 
found by multiplying the net area of the plunger by the length of the stroke 
in feet (in direct-connected engines the average length of stroke); then by the 
number of single strokes per minute; and the product of these three by the 
constant 74.8; finally correcting for the percentage of leakage of the pump. 

(d) Leakage of Pump. The percentage of leakage is the percentage borne by 
the quantity of leakage found on the leakage trial, to the quantity of water 
discharged on the duty run determined from plunger displacement. 

(e) Friction. The percentage of total friction in a reciprocating pump is the 
percentage borne by the friction horse power to the indicated horse power of 
the steam cylinders. 

(/) Miscellaneous. For the calculation of other results pertaining specially to 
the performance of the steam end of a reciprocating pump, reference may be 
made to the Steam Engine Code. 



8. DATA AND RESULTS. 

The data and results should be reported in accordance with the form 
given herewith, adding lines for data not provided for, or omitting those 
not required, as may conform to the object in view. 

In the case of a pumping engine of the reciprocating class for which 
a record of the complete performance is desired, the additional engine 
data and results given in the Complete Form of the Steam Engine Code 
may supplement those here given. 

* If there is a material difference in velocity of the water at the points where the gauges are attached, 
a correction should be made for the corresponding difference in " velocity head." 



950 STEAM POWER PLANT ENGINEERING 



DATA AND RESULTS OF DUTY TRIAL OF STEAM PUMPING 

MACHINERY. 

CODE OF 1912. 

(1) Test of pump, located at 

to determine , conducted by 

(2) Type of machinery 

(3) Rated capacity in gallons per 24 hours 

(4) Type of boiler 

(5) Type of auxiliaries 

(6) Dimensions of engine or turbine 

(7) Dimensions of pump 

(8) Dimensions of boilers 

(9) Dimensions of auxiliaries 

(10) Dimensions of condenser 

(11) Date 

(12) Duration hr. 

Average Pressures and Temperatures. 

(13) Steam pressure at boiler by gauge lb. per sq. in. 

(14) Steam-pipe pressure near throttle, by gauge lb. per sq. in. 

(15) Barometric pressure of atmosphere in inches of mercury in. 

(16) Pressure in receiver by gauge lb. per sq. in. 

(17) Vacuum in condenser in inches of mercury in. 

(18) Pressure in force main by gauge .lb. 

(19) Pressure in suction main by gauge lb. 

(20) Vertical distance between centers of two gauges ft. 

(21) Total head, expressed in lb. pressure .lb. 

(22) Total head, expressed in feet ft. 

(23) Temperature of main supply of feed water to boilers deg. 

(24) Temperature of additional supplies of feed water deg. 

(25) Temperature of air in engine room deg. 

Total Quantities 

(26) Water fed to boilers from main source of supply lb. 

(27) Water fed from additional supplies lb. 

(28) Total water fed to boilers from all sources lb. 

(29) Moisture in steam or superheating near throttle per cent or deg. 

(30) Factor of correction for quality of steam, dry steam being unity 

(31) Total dry steam consumed for all purposes lb. 

(32) Total leakage of plungers and valves gal. 

(33) Total number of gallons of water discharged: 

(a) By plunger displacement, uncorrected gal. 

(6) By plunger displacement, corrected for leakage gal. 

(c) By weir or other measurement of discharge gal. 

(34) Total weight of water discharged: 

(a) By plunger displacement, corrected lb. 

(6) By measurement of discharge lb. 



APPENDIX F 951 

Hourly Quantities. 

(35) Water fed from main source of supply lb. 

(36) Water fed from additional supplies lb. 

(37) Total water fed to boilers per hour lb. 

(38) Total dry steam consumed per hour lb. 

(39) Loss of steam and water per hour due to drips from main steam pipes and 

to leakage of plant lb. 

(40) Net dry steam consumed per hour lb. 

(41) Dry steam consumed per hour: 

(a) By main engine or turbine lb. 

(6) By auxiliaries lb. 

(42) Water discharged per hour : 

(a) By plunger displacement, corrected lb. 

(6) By measurement of discharge lb. 

Heat Data. 

(43) Heat units per lb. of dry steam based on temperature, Line 23 B.t.u. 

(44) Heat units per lb. of dry steam based on temperature, Line 24 B.t.u. 

(45) Heat units consumed per hour based on main supply of feed B.t.u. 

(46) Heat units consumed per hour based on additional supplies of feed B.t.u. 

(47) Total heat units consumed per hour for all purposes B.t.u. 

(48) Loss of heat per hour due to leakage of plant, drips, etc B.t.u. 

(49) Net heat units consumed per hour B.t.u. 

(50) Heat units consumed per hour : 

(a) By engine or turbine alone B.t.u. 

(b) By auxiliaries B.t.u. 

Indicator Diagrams. 

(51) Mean effective pressure lb. 

Speed and Stroke. 

(52) Revolutions per minute rev. 

(53) Number of single strokes per minute strokes. 

(54) Average length of stroke ft. 

Power. 

(55) Indicated horse power developed: 

1st cylinder i.h.p. 

2d cylinder .i.h.p. 

Whole engine i.h.p. 

(56) Water horse power h.p. 

(57) Friction h.p. (Line 55 — Line 56) r.h.p. 

(58) Percentage of i.h.p. lost in friction .per cent. 

Duty. 

(59) Duty per million heat units ft.-lb. 

(60) Duty per thousand lb. of steam ft.-lb. 



952 STEAM POWER PLANT ENGINEERING 

Work Done per Heat Unit. 

(61) Ft.-lb. of work per B.t.u. (Line 59 -^ 1,000,000) ft.-lb. 

Capacity. 

(62) Number of gal. of water pumped in 24 hr. : 

(a) By plunger displacement, corrected for leakage gal. 

(6) By weir or other measurement gal. 

(63) Gallons of water per minute: 

(a) By plunger displacement, corrected for leakage gal. 

(6) By weir or other measurement gal. 

Economy Results, Steam End. 

(64) Heat units consumed per i.h.p. per hour:* 

(a) By engine and auxiliaries . . B.t.u. 

(6) By engine alone B.t.u. 

(c) By auxiliaries B.t.u. 

(65) Dry steam consumed per i.h.p. per hour: 

(a) By engine and auxiliaries lb. 

(b) By engine alone lb. 

(c) By auxiliaries lb. 

* The i.h.p. on which the economy results are based is that of the main engine given in Line 54. 






APPENDIX G. 

RULES FOR CONDUCTING TESTS OF COMPLETE STEAM PUMPING 
MACHINERY PLANTS.* 

A.S.M.E. Code of 1912. 
1. OBJECT AND PREPARATIONS. 

This code applies to a commercial test of a complete steam pumping 
machinery plant, having for an object the determination of the fuel 
cost of pumping a given quantity of water for a day's run of 24 hours. 
For tests of the component parts of the plant, rules may be found in 
the respective codes applying thereto. 

Read the general instructions given in Appendix A. Take the 
dimensions, note the physical conditions, examine for leakages, install 
the testing appliances, etc., as there pointed out and prepare for the 
test accordingly. 

In a reciprocating pump, ascertain the quantity of water leakage or 
slip past the plungers, and that of the pump valves. 

2. FUEL. 

Determine the character of the fuel to conform with the object in 
view. To obtain maximum economy or capacity, the fuel should be 
some kind of coal that is regarded as a standard, as noted in the Boiler 
Code. 

3. APPARATUS AND INSTRUMENTS. 

The apparatus and instruments required for a simple performance 
test of a complete pumping machinery plant are : 

(a) Platform scales for weighing coal and ashes. 
(6) A coal calorimeter. 

(c) A tachometer or other speed-measuring apparatus. 

(d) A pressure gauge attached to the force main. 

(e) For rotary pumps, a weir or other means for measuring the quantity of water 

pumped. 
(/) Stroke scales, for direct-acting pumps. 

If the test also involves the determination of boiler performance 
and pumping machinery performance, additional apparatus and in- 

* Preliminary report of the Committee on Power Tests. (Jour. A.S.M.E., Nov., 
1912.) 

953 



954 STEAM POWER PLANT ENGINEERING 

struments are required, as pointed out in the respective codes referring 
thereto. 

4. OPERATING CONDITIONS. 

Determine what the operating conditions should be to conform to 
the object in view, and see that they prevail throughout the trial. 

5. DURATION. 

The duration of a test of a complete pumping machinery plant should 
be not less than one day of 24 hours. 

In cases where the machinery is in operation only a part of the calen- 
dar day, the coal consumption used in computing the results should 
be that of the entire 24 hours. 

6. MISCELLANEOUS. 

The methods of starting and stopping a test of a complete pumping 
plant, sampling and drying coal, determining ashes and refuse, and 
ascertaining the calorific value and chemical analysis of the coal, are 
the same as those described in the Code for Complete Steam Power 
Plants. 

7. CALCULATION OF RESULTS. 

For the calculation of water horse power, duty, capacity, and leak- 
age of pump, reference may be made to the explanations given in the 
Pumping Engine Code. 

8. RECORDS, DATA, AND RESULTS. 

The records should be made as explained in Appendix A, and the 
tabular summary of data and results reported in accordance with the 
form given herewith, adding lines not provided for, or omitting those 
not required. 

DATA AND RESULTS OF TEST OF STEAM PUMPING MACHINERY 

PLANT. 

CODE OF 1912. 



(1) Test of plant located at ... . 

to determine conducted by . 

(2) Type of machinery 

(3) Rated capacity in gallons per 24 hrs 

(4) Type of boilers 

(5) Type of auxiliaries 

(6) Dimensions of machinery 

(7) Dimensions of boilers 






APPENDIX G 955 

(8) Dimensions of auxiliaries 

(9) Dimensions of condenser 

(10) Date 

(11) Duration hr. 

(12) Length of time machinery was in motion with throttle open hr. 

(13) Length of time machinery was running at normal speed hr. 

Average Pressures and Temperatures. 

(14) Steam pressure at boiler by gauge lb. per sq. in. 

(15) Pressure in force main by gauge lb. 

(16) Pressure in suction main by gauge in. 

(17) Vertical distance between centers of two gauges ft. 

(18) Total head expressed in lb. pressure lb. 

(19) Total head expressed in feet ft. 

Speed and Stroke. 

(20) Revolutions per minute rev. 

(21) Number of single strokes per minute strokes. 

(22) Average length of stroke ft. 

Total Quantities. 

(23) Total coal as fired lb. 

(24) Moisture in coal per cent. 

(25) Total dry coal consumed lb. 

(26) Ash and refuse . lb. 

(27) Percentage of ash and refuse to dry coal per cent. 

(28) Calorific value per lb. of dry coal by calorimeter test B.t.u. 

(29) Cost of coal per ton of lb dollars. 

(30) Total leakage of plungers and valves gal. 

(31) Total number of gals, of water discharged: 

(a) By plunger displacement, uncorrected gal. 

(6) By plunger displacement, corrected for leakage gal. 

(c) By weir or other measurement gal. 

Hourly Quantities. 

(32) Dry coal consumed per hour, based on duration of running period (Line 25 

i - -7- Line 12) lb . 

Power. 

(33) Water horse power h.p. 

Economy Results. 

(34) Dry coal consumed per water h.p. per hour lb. 

(35) Cost of coal per water h.p. per hour '. cents. 

(36) Duty per 100 lb. of dry coal 

Capacity. 

(37) Number of gal. of water pumped in 24 hr. : 

(a) By plunger displacement, corrected for leakage gal. 

(b) By weir or other measurement gal. 

(38) Number of gal. of water pumped per minute: 

(a) By plunger displacement, corrected for leakage gal. 

(b) By weir or other measurement gal. 



APPENDIX H. 

RULES FOR CONDUCTING TESTS OF STEAM-DRIVEN COMPRESSORS, 
BLOWERS AND FANS.* 

A.S.M.E. Code of 1912. 

1. OBJECT AND PREPARATIONS. 

2. APPARATUS AND INSTRUMENTS. 

3. OPERATING CONDITIONS. 

4. DURATION. 

5. STARTING AND STOPPING. 

6. RECORDS. 

The directions pertaining to the above divisions of the subject are 
almost identical with those given in the Pumping Machinery Code 
under the same headings, the only difference being that the work done 
by the steam end is expended in moving a volume of air, instead of a 
body of water. 

The quantity of air discharged should be measured by a gasometer, 
or by delivery into tanks of known capacity. Where these means are 
not available the pitot tube should be used, or some other means should 
be employed which is subject to calibration. 

If the air end is of the reciprocating type, indicator diagrams should 
be regularly taken from this end as well as from the steam end. 

7. CALCULATION OF RESULTS. 

The rules pertaining to dry steam, heat consumption, and indicated 
horse power of the steam end, are identically the same as those given 
in the Steam Engine Code, and reference may be made to that code for 
the necessary directions in these particulars. 

(a) Air Horse Power. The gross work done at the air end of a reciprocating 
machine, expressed in horse power, is found by multiplying together the net 
area of the air piston in square inches, the mean effective air pressure in 
pounds per square inch as determined from indicator diagrams, the length of 

* Preliminary report of Committee on Power Tests. (Jour. A.S.M.E., Nov., 1912.) 
In the case of air machinery driven by some other prime mover than a steam engine 
or turbine, the code may be modified to meet the particular requirements. 

956 









APPENDIX H 957 

the stroke in feet, and the number of single strokes per minute; and dividing 
their product by 33,000. 

The net work at the air end of either reciprocating or rotary machines, 
expressed in foot-pounds per minute, is found by multiplying the corrected 
volume of the compressed air in cubic feet discharged into the main delivery 
pipe per minute, by the impact or total pressure in pounds per square feet 
and by the hyperbolic logarithm of the ratio of the total pressure to the 
atmospheric pressure (all pressures being absolute pressures). The net air 
horse power is found by dividing the product by 33,000. The corrected 
volume of the compressed air may be found by multiplying the sectional 
area of the delivery main in square feet by the mean velocity in feet per 
minute as determined by pitot tube or other measurement, and reducing the 
result to atmospheric temoerature by multiplying by the proportion 

460 + t 
460 + T ' 

in which t is the temperature of the air supplied to the machine and T the 
temperature of the air in the delivery main. 

(b) Capacity. The capacity is the number of cubic feet of air discharged 
through the delivery main per minute, as determined by gasometer, tank or 
other mode of measurement, reduced to the equivalent free air at the atmos- 
pheric temperature and pressure. The correction for pressure is made by 

multiplying by the proportion ■=?, in which Pi is the atmospheric pressure 

"x 

and P 2 the total pressure in the main (absolute pressures), and the correction 
for temperature as above. 

The capacity may also be expressed in the number of cubic feet of com- 
pressed air discharged per minute at a given pressure above the atmosphere 
reduced to the atmospheric temperature. 

(c) Miscellaneous. For methods of calculating results pertaining especially 
to the performance of the steam-end of a reciprocating air-pumping machine, 
reference may be made to the Steam Engine Code. 

The "efficiency of compression" in a reciprocating machine is determined 
by first ascertaining the net work at the air end given above under the head- 
ing, " (a) Air Horse Power," and then dividing the net work thus found by 
the gross work given under the same heading. 

The "mechanical efficiency" of a reciprocating machine is determined by 
dividing the gross air horse power at the air end by the indicated horse power 
at the steam end, or by the horse power delivered by the belt or motor in the 
case of other means of driving. 



8. DATA AND RESULTS. 

The data and results should be reported in accordance with the form 
given herewith, adding lines for data not provided for and omitting 
those not required, as may conform with the object in view. 

In the case of an air-pumping machine of the reciprocating class for 
which a record of the complete performance is desired, the additional 



958 STEAM POWER PLANT ENGINEERING 

engine data and results given in the Complete Form of the Steam 
Engine Code may supplement those here given: 



DATA AND RESULTS OF TEST OF AIR MACHINERY. 
CODE OF 1912. 

(1) Test of located at 

to determine conducted by 

(2) Type of machinery 

(3) Rated capacity in cu. ft. of free air per minute 

(4) Rated capacity in cu. ft. of air discharged per minute at 100 lb. per sq. in. 

above atmosphere, reduced to the atmospheric temperature cu. ft. 

(5) Type of boilers 

(6) Type of auxiliaries 

(7) Dimensions of engine or turbine at steam end 

(8) Dimensions of cylinders or blowers at air end 

(9) Dimensions of boilers 

(10) Dimensions of auxiliaries 

(11) Dimensions of condenser 

(12) Date 

(13) Duration hr. 



Average Pressures and Temperatures. 

(14) Steam pressure at boiler by gauge lb. per sq. in 

(15) Steam-pipe pressure near throttle, by gauge lb. per sq. in 

(16) Barometric pressure of atmosphere in in. of mercury in 

(17) Pressure in receiver by gauge ~ lb. per sq. in 

(18) Vacuum in condenser in in. of mercury in 

(19) Pressure in delivery main by gauge (impact pressure) lb 

(20) Total head, expressed in ft ft 

(21) Temperature of main supply of feed water to boilers .deg 

(22) Temperature of additional supplies of feed water deg 

(23) Temperature of air in engine room or air supplied to machine deg 

(24) Temperature by wet-bulb thermometer deg 

(25) Temperature of air in delivery main deg 



Total Quantities. 

(26) Water fed to boilers from main source of supply lb. 

(27) Water fed from additional supplies lb. 

(28) Total water fed to boilers from all sources lb. 

(29) Moisture in steam or superheating near throttle per cent or deg. 

(30) Factor of correction for quality of steam, dry steam being unity 

(31) Total dry steam consumed for all purposes lb. 

(32) Total cu. ft. of compressed air delivered as measured cu. ft. 

(33) Total cu. ft. of compressed air delivered reduced to atmospheric tempera- 

ture and pressure lb. 

(34) Total weight of air delivered lb. 



APPENDIX H 959 



Hourly Quantities. 

(35) Water fed from main source of supply lb. 

(36) Water fed from additional supplies lb. 

(37) Total water fed to boilers per hour lb. 

(38) Total dry steam consumed per hour lb. 

(39) Loss of steam and water per hour due to drips from main steam pipes and 

to leakage of plant lb. 

(40) Net dry steam consumed per hour lb. 

(41) Dry steam consumed per hour: 

(a) By main engine or turbine lb. 

(b) By auxiliaries lb. 

(42) Cu. ft. of compressed air delivered per hour as measured cu. ft. 

(43) Cu. ft. of compressed air delivered per hour reduced to atmospheric tem- 

perature cu. ft. 

(44) Cu. ft. of compressed air delivered per hour reduced to atmospheric tem- 

perature and pressure , . . cu. ft. 

(45) Weight of air delivered per hour lb. 



Heat Data. 

(46) Heat units per lb. of dry steam based on temperature Line 21 B.t.u. 

(47) Heat units per lb. of dry steam based on temperature Line 22 B.t.u. 

(48) Heat units consumed per hour based on main supply of feed B.t.u. 

(49) Heat units consumed per hour based on additional supplies of feed B.t.u. 

(50) Total heat units consumed per hour for all purposes B.t.u. 

(51) Loss of heat per hour due to leakage of plant, drips, etc B.t.u. 

(52) " Net heat units consumed per hour B.t.u. 

(53) Heat units consumed per hour: 

(a) By engine or turbine alone B.t.u. 

(6) By auxiliaries B.t.u. 



Indicator Diagrams. 

(54) Mean effective pressure in steam cylinders lb. 

(55) Mean effective pressure in air cylinders lb. 



Speed and Stroke. 

(56) Revolutions per minute rev. 

(57) Number of single strokes per minute strokes. 



Power. 

(58) Indicated horse power developed at steam end of reciprocating machine . . . i.h.p. 

(59) Gross air horse power as indicated in air cylinders of reciprocating 

machine air h.p. 

(60) Net air horse power as computed from Line 43 air h.p. 

(61) Friction of reciprocating machine (Line 58 — Line 59) fr.h.p. 

(62) Percentage of i.h.p. lost in friction of machine per cent. 



960 STEAM POWER PLANT ENGINEERING 

Economy Results, Steam End of Engine-driven Machines. 

(63) Heat units consumed per i.h.p. per hour: 

(a) By engine and auxiliaries B.t.u. 

(6) By engine alone B.t.u. 

(c) By auxiliaries B.t.u. 

(64) Dry steam consumed per i.h.p. per hour: * 

(a) By engine and auxiliaries lb. 

(6) By engine alone , lb. 

(c) By auxiliaries lb. 

Economy Results, Air Delivered. 

(65) Heat units consumed per hour per net air h.p. of Line 60: 

(a) By engine or turbine and auxiliaries B.t.u. 

(6) By engine or turbine alone B.t.u. 

(c) By auxiliaries B.t.u. 

(66) Dry steam consumed per hour per net h.p. of line 60: 

(a) By engine or turbine and auxiliaries lb. 

(b) By engine or turbine alone lb. 

(c) By auxiliaries lb. 

Efficiency Results. 

(67) Thermal efficiency ratio for engine alone: 

(a) Per i.h.p., steam end (2545 -r- Line 636) per cent. 

(b) Per net air h.p., air delivery (2545 -r- Line 65&) per cent. 

Work Done per Heat Unit. 

(68) Ft.-lb. of net work per B.t.u. consumed by engine or turbine and auxilia- 

ries (1,980,000 -r- Line 65a) ft.-lb. 

Capacity. 

(69) Cu. ft. of compressed air delivered per minute as measured cu. ft. 

(70) Cu. ft. of compressed air delivered per minute, reduced to atmospheric 

temperature cu. ft. 

(71) Cu. ft. of compressed air delivered per minute at 100 lb. pressure, reduced 

to atmospheric temperature cu. ft. 

(72) Cu. ft. of compressed air delivered per minute, reduced to atmospheric 

temperature and pressure (free air) cu. ft. 

Miscellaneous Results. 
Steam-driven Reciprocating Machine. 

(73) Efficiency of compression (Line 60 4- Line 59 X 100) per cent. 

(74) Mechanical efficiency of machine (Line 59 + Line 58 X 100) per cent. 

(75) Volumetric efficiency (Line 72 -f- 1st Compr. Displ. X 100) per cent. 

* The i. h.p. on which these economy results are based is that of the main engine given in Line 58. 

Note: In the case of air compressors having more than one stage and in those having intercoolers 
additional data should be given covering pressures and temperatures in the different stages, the quan- 
tity of water used for cooling and temperatures of the air and water entering and leaving the cooler. 



APPENDIX I 



961 



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962 



STEAM POWER PLANT ENGINEERING 





Density 
Weight per 
Cubic Foot, 

Pounds. 


J- 


03330 
03569 
03732 


03806 
04980 
0614 


0728 
0841 
0953 


1065 
1175 
1285 


1394 
1503 
1612 


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1829 
1937 


2044 
2151 

2258 




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APPENDIX I 



963 



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APPENDIX J. 

EQUIVALENT VALUES OF ELECTRICAL AND MECHANICAL UNITS. 



1 Myriawatt = 


1 Kilowatt = 


10 kilowatts 


0.1 myriawatt 


10,000 watts 


1,000 watts 


13.41 horsepower 


1.341 horse power 


13.597 cheval-vapeur 


1.3597 cheval-vapeur 


13.597 pferde-kraft 


1.3597 pferde-kraft 


26,552,000 foot pounds per hour 


2,655,200 foot pounds per hour 


8,605,000 gram calories per hour 


860,500 gram calories per hour 


3,670,000 kilogram meters per hour 


367,000 kilogram meters per hour 


34,150 B.t.u. per hour 


3,415 B.t.u. per hour 


1.02 boiler horse power 


0.102 boiler horse power 



Horse Power = 
745.7 watts 
0.7457 kilowatt 
0.07457 myriawatt 
1.0139 cheval-vapeur 
1.0139 pferde-kraft 
33,000 foot pounds per minute 
641,700 gram calories per hour 
273,743 kilogram meters per hour 
2,547 B.t.u. per hour 



1 Cheval-Vapeur or Pferde-Kraft 
75 kilogram meters per second 
0.07354 myriawatt 
0.7357 kilowatt 
0.9863 horse power 
32,550 foot pounds per minute 
632,900 gram calories per hour 
2,512 B.t.u. per hour 



1 Joule = 

1 watt second 
0.10197 kilogram meter 
0.73756 foot pound 
0.239 gram calorie 
0.0009486 B.t.u. 



1 Foot Pound = 

1.3558 joules 
0.13826 kilogram meter 
0.001286 B.t.u. 
0.03241 gram calorie 
0.000000505 horse-power hour 



1 B.t.u. = 

1,054 watt seconds 
777.5 foot pounds 
107.5 kilogram meters 

0.0003927 horse-power hour 



1 Kilogram- Meter = 
7.233 foot pounds 
9.806 joules 
2.344 gram calories 
0.0093 kilogram meter 



964 



APPENDIX K. 

MISCELLANEOUS CONVERSION FACTORS. 



1 Pound per Square Inch = 

2.0355 inches of mercury at 32° F. 
2.0416 inches of mercury at 62° F. 
2.309 feet of water at 62° F. 
0.07031 kilogram per square centi- 
meter 
0.06804 atmosphere 
51.7 millimeters of mercury at 

32° F. 

1 Foot of Water at 62° F. = 
0.433 pound per square inch 
62.355 pounds per square foot 
0.883 inch of mercury at 62° F. 
821.2 feet of air at 62° F. and 
barometer 29.92 

1 Inch of Water 62° F. = 

0.0361 pound per square inch 
5.196 pounds per square foot 
0.5776 ounce per square inch 
0.0736 inch of mercury at 62° F. 
68.44 feet of air at 62° F. and 
barometer 29.92 

1 Foot of Air at 32° F. and Barometer 
29.92 = 
0.0761 pound per square foot 
0.0146 inch of water at 62° F. 

1 Inch of Mercury at 62° F. = 
0.4912 pound per square inch 
1.132 feet of water at 62° F. 
13.58 inches of water at 62° F. 



1 Atmosphere = 

760.0 millimeters of mercury at 
32° F. 
14.7 pounds per square inch 
29.921 inches of mercury at 32° F. 
2,116.0 pounds per square foot 

1.033 kilograms per square centi- 
meter 

1 Millimeter = 0.03937 inch 

1 Centimeter = 0.3937 inch 

1 Meter = 39.37 inches 

1 Meter = 3.2808 feet 

1 Square Meter = 10.764 square feet 

1 Liter = 

61.023 cubic inches 
0.264 U. S. gallons 

1 Gram = 

1 cubic centimeter of distilled 

water 
15.43 grains troy 
0.0353 ounce 

1 Kilogram = 

2.20462 pounds avoirdupois 



965 



APPENDIX L. 

TOTAL HEAT-ENTROPY DIAGRAM. 
(mollier diagram.) 

The steam tables give values of the simultaneous physical properties 
of steam, such as pressure, entropy, temperature, etc. When certain 
of these properties are known the remainder can be obtained from the 
tables. The simultaneous properties can also be shown by means of a 
diagram each point on which represents steam in a perfectly definite 
condition. 

Fig. 608 gives a skeleton outline of such a diagram and Fig. 609 a 
reduced reproduction of the complete chart as constructed by Marks 



I 

a 



- — *X 4 \/ 




£ \% 

Xj^r__ _^-<*iabatic_ _ 

Xl j ^\ 
Pi 


x 3 

/s X^ 

x 2 



^ . 



B.T.TT.per Pound of Steam 
Fig. 608. 



and Davis. Referring to Fig. 608, abscissas represent the heat con- 
tents or B.t.u. per pound of steam and ordinates represent the total 
entropy. Vertical lines then represent lines of constant heat content, 
and horizontal lines constant entropy. P\P\ and P2P2 represent lines 
of constant pressure and X1X1 and X 2 X 2 lines of constant quality. Evi- 
dently any point in the chart represents a fixed condition of heat con- 
tent, pressure, quality and entropy as determined by its location with 
respect to the different lines. Thus point 1 represents a pressure Pi as 

966 



1300 



1250 



1200 



1150 



1100 




105H 



1000 



El|OlT. 

Ft. lbs. B.T.U 
of Work 




1.78 1-82 



APPENDIX L 967 

determined by the numerical value of line P1P1, quality Xi by its loca- 
tion on line XiX h entropy ni by its projection Ni on the Y axis, and 
heat content Hi by its projection on the X axis. 

The principal advantages of a total heat-entropy diagram over the 
tables are that they give the properties of wet and superheated steam 
and offer a simple means of solving many problems without calculations. 
For example, the chart offers a ready solution of problems involving 

(a) Adiabatic expansion. 

(b) Throttling. 

(c) Expansion with frictional resistances. 

(a) Adiabatic Expansion: From thermodynamics we know that dur- 
ing an adiabatic change the entropy is constant; thus, in expanding 
from pressure Pi and condition represented in point 1 to a lower pressure 
P 2 it is only necessary to find the intersection 2 of a horizontal line from 
point 1 with line P2P2. The various properties corresponding to point 
2 can be read directly from the diagram. 

The line 1-2 = HiH 2 represents the difference in heat content following 
adiabatic expansion from pressure Pi and condition 1 to pressure P, or 

line HiH 2 = Hi — H 2 = x&i + qi — x 2 r 2 — q 2 . 

The quality x 2 is read directly from the intersection of line 1-2 with 
the constant quality line X 2 X 2 . 

The entropy n 2 , of course, remains the same. 

From equation (129), p. 389, we find that the velocity due to adia- 
batic expansion is 

V= 223.8 VHi-H 2 . 

Marks and Davis have added along the margin of the diagram (Fig. 
609) a scale of velocity so that V may be ascertained by laying off the 
length HiH 2 on the scale. 

Example: Steam at 120 pounds absolute, quality 0.98, expands adia- 
batically to a back pressure of 2 pounds absolute. Find the quality 
and heat content at the lower pressure. 

From Fig. 609 we locate Pi at the intersection of pressure curve 120 
and quality curve 0.98. The corresponding values of Hi and n x are 
found by interpolation to be 1174.7 and 1.564 respectively. Follow 
vertical line 1.564 until it intersects pressure line " 2 ". The correspond- 
ing values of H 2 and x 2 are found to be 910 and 0.797 respectively. 
The vertical intercept between the two pressure lines laid off on the 
velocity diagram gives V = 3640 feet per second. 

Supposing the steam to be superheated 200 degrees instead of being 
wet, find the quality and heat content at the end of expansion. 



968 STEAM POWER PLANT ENGINEERING 

Locate Pi at the intersection of pressure curve 120 and superheat 
curve 200. The corresponding values of Hi and n\ are found to be 
1295 and 1.703 respectively. Follow vertical line 1.703 until it inter- 
sects pressure line "2". The corresponding values of H 2 and x^ are 
found to be 990 and 877 respectively. 

(6) Throttling: If steam expands through a small orifice without the 
addition or abstraction of heat and is brought finally to its initial ve- 
locity its total heat will be unchanged. This process is called throttling 
and occurs when steam passes through a reducing valve. Vertical lines 
in Fig. 608 are lines of constant total heat and consequently show the 
changes in the condition of steam which result from throttling. Thus 
in throttling steam from pressure Pi, Fig. 608, to P 2 it is only necessary 
to find the intersection, 4, of a vertical line from point 1 with line P2P2. 

Example: Steam at 200 pounds pressure and quality 0.96 passes 
through a reducing valve and its pressure is lowered to 15 pounds. 
Find its quality at the lower pressure. The intersection of pressure 
line 190, Fig. 609, with quality line 0.96 gives Hi = 1165. Follow hori- 
zontal line 1165 until it intersects pressure line 15. The corresponding 
value for x? is found to be 30, that is, the steam is superheated 30 de- 
grees. 

To what pressure must the steam be reduced in order that it may be 
dry and saturated? Follow horizontal line 1165 until it intersects the 
saturation curve. The corresponding pressure is found to be 30 pounds. 

(c) Expansion Involving Frictional Resistances: As steam expands 
in the nozzle of a turbine or passes between the vanes it experiences 
frictional resistances which cause it to give up less energy than it would 
under ideal conditions. The work of friction causes the entropy of the 
steam at its lowest temperature to be greater than it would be if adia- 
batic expansion occurred and serves to increase its dryness fraction. 

If y one hundredths of the heat Hi — H 2 (given up in adiabatic ex- 
pansion) is lost due to friction, the heat available for useful work is 

(1 - y) (ft - ft), 
the resulting velocity of the jet is 



V=22S.sV(l-y)(Hi-H 2 ), 

and the increase in quality of the exhaust steam is 

y{Hi-H 2 ) ^ 
r 2 

These equations may be readily solved by means of the diagram. 
Referring to Fig. 608 line 1-3 represents an expansion from pressure 
Pi to pressure P 2 with frictional resistances. 






APPENDIX L 969 

From the diagram 

N2 H 3 H 2 

y 



12 H 2 H 1 

(1 - y) (Hi - H 2 ) = line 1 N = HiH z . 

Increase in quality = = distance 2-3 between X 2 X 2 and 

T 2 

XzX$. 

Increase in entropy = N 3 = N1N3. 

Example: Steam at 160 pounds absolute initial pressure, quality 0.97, 
expands through a nozzle to a back pressure of 2 pounds absolute. If 
15 per cent of the heat energy is lost in friction, find the quality of the 
steam at the lower pressure and the velocity of the jet. 

From Fig. 609 we locate Pi at the intersection of pressure curve 160 
and quality curve 0.97. The corresponding value of Hi is 1170. Fol- 
low line 1170 vertically until it intersects pressure line "2". From the 
diagram we find for adiabatic expansion H 2 = 910 x 2 = 0.797. But the 
friction increases the heat content at the end of expansion an amount 
0.15 X #i - H 2 = 0.15 (1170 - 910) = 39, so that the final heat con- 
tent = 910 + 39 = 949. 

Follow pressure line "2" until it intersects heat line 949. The qual- 
ity X2 is found to be 0.836 and the entropy ?h = 1.632. From the 
velocity scale we find V = 3320 feet per second for Hi — H 2 = (1170 
- 949). 



INDEX 



Acetylene, properties of, 33. 

Acidity, tests for, in oils, 756. 

Acme bucket trap, 659. 

Acton atmospheric relief valve, 747. 

Air-back, swinging, 176. 

Air chambers, 594. 

Air-cooled surface condensers, 499. 

Air lift, 642. 

Air, properties of, 524. 

Air pumps, 618-628. 

Alberger rotative, 627. 

Dean, 618. 

Edwards, 622. 

Hewes and Phillips, 620. 

Leblanc, 624. 

Mullan, 622. 

size of dry, 627. 

size of wet, 619, 621, 626. 

Wheeler "Rotrex," 623. 
Air required for combustion, 36. 
Air required for fuel oils, 69. 
Air required for operating air lift, 642. 
Air requirements for smokeless combus- 
tion, 153. 
Air spaces, grate bars, 143. 
Air thermometers, recording, 793. 
Air vs. steam as an oil atomizer, 79. 
Air, weight of per pound of carbon, 42. 
Alarm, high- and low-water, 149. 
Alberger barometric condenser, 479. 

cooling tower, 521. 

rotative dry-air pumps, 627. 
AUis-Chalmers steam turbine, 435. 
American underfeed stoker, 183. 
Analysis of boiler scales, 540. 

of flue gases, 47, 801. 

of fuel oils, 67. 

of typical American coals, 16-26. 

of waters for boiler feeding, 540. 
Anchors, pipe, 697. 

Anderson automatic non-return valve, 
738. 



Anderson feed-water purifying system, 
551. 

triple-duty emergency valve, 739. 
Animal fats and oils, 750. 
Anthracite coals, 18. 
Aqueous vapor, effect of, on degree of 
vacuum, 473. 

pressure of, 465. 
Arch bars, furnace, 101. 
Arches, deflecting, 159. 
Armour Glue Works, vacuum ash system 

at, 238. 
Armour Institute, brick chimney at, 266. 
Ash bins, 223. 

Ash conveyor, vacuum system, 238. 
Ash-handling systems, 222-240. 
Ash, influence of, on fuel value of dry 

coal, 57. 
A.S.M.E. rules, boiler trials, 919. 

code of 1912. 

compressors, blowers and fans, 956. 

power plants, 943. 

pumping engines, 947. 

reciprocating engines, 927. 

turbines and turbo-generators, 939. 
Atmospheric heaters, 554. 
Atmospheric relief valves, 747. 
Atmospheric surface lubrication, 757. 
Atomization of oil, 79. 
Augmenter, Parsons vacuum, 488. 
Aurora and Elgin Interurban Ry., coal- 
handling system, 234. 
Austin steam separator, 648. 
Automatic cut-off vs. throttling engines, 

364. 
Automatic injectors, 613. 
Automatic non-return valves, 738. 
Automatic stage valve, Curtis turbine, 

415. 
Automatic temperature control, 726. 
Auxiliaries, measurement of steam used 
by, 510. 



971 



972 



INDEX 



Auxiliaries, power consumption of con- 
denser, 510. 

Babcock & Wilcox boilers, 104. 

chain grate, 173. 

superheater, 209. 
Back connection, return tubular boiler, 

101. 
Back pressure on engines, 334. 
Back-pressure valves, 746. 
Baffle-plate steam separator, 648. 
Bagasse as fuel, 26. 
Balanced-draft system, 306. 
Baragwanath feed-water heater, 563. 

siphon condenser, 475. 

surface condenser, 482. 
Barnard- Wheeler cooling tower, 520. 
Barometric condensers, 477. 
Baum oil separator, 652. 
Baume, gravity, 755. 
Bearings, lubrication of, 757. 
Belliss engines, tests of, with super- 
heated steam, 368. 
Belt conveyors, 232. 
Bends, pipe, 695. 
Bigelow-Hornsby boiler, 113. 
Binary-vapor engines, 379. 
Bituminous coals, 20-23. 
Bladeless turbine, Tesla, 458. 
Blades, arrangement of, in steam tur- 
bines, 392. 
Blake jet condenser, 469. 
Blake pulverizer and feeder, 62. 
Blast-furnace gas, properties of, 87. 
Bleeder type of turbine, 439. 
Blonck boiler efficiency meter (re/.), 798. 
Bloomsburg steam jet, 286. 
Blowers, fan, 289. 

steam jet, 286. 

tests of, 292. 
Blow-off piping, South Side Elevated 

R.R., 742. 
Blow-offs, 145. 
Blow-off tank, 145. 
Blow-off valve, 741. 
Boiler cleaners, 150. 
Boiler compounds, 543. 
Boiler-feed pumps (see Pumps). 
Boiler room area, 110. 
Boiler tests, A.S.M.E. code, 916. 



Boiler tubes, size of standard, 681. 
Boilers, 89-152. 

Babcock & Wilcox, 104. 

Bigelow-Hornsby, 113. 

capacity of, 130, 135. 

classification of, 89. 

cost of, 141. 

efficiency of, 123. 

efficiency of perfect, 125. 

fire-box, 91. 

Fitzgibbons, 92. 

forced capacity of, 134. 

furnaces for, 153-173. 

grates for, 142, 178. 

heating surface of, 118. 

Heine, 106. 

horizontal return tubular, 96. 

horse power of, 120. 

inspector's report (1911), Hartford 
Boiler Insurance Co., 541. 

Lyons, 102. 

Manning vertical, 91. 

Parker boiler, 107. 

performance of, 126. 

Robb-Mumford, 95. 

Scotch marine, 93. 

Sederholm, 102. 

selection of type, 141. 

settings for, 90-180. 

specifications for, 865. 

Stirling, 110.- 

temperature drop in, 140. 

tests of (see Tests). 

vertical tubular, 90. 

Wickes, 107. 
Bone, surface combustion, 86. 
Boston Elevated, cost of operation, 846. 
Bourdon pressure gauge, 792. 
Brake horse power, 323. 
Branch fuel oil burner, 74 
Brass pipes, 678. 
Breeching, 280. 
Brick chimneys, 265. 
Bridge- wall, hollow, 168. 
Bristol recording air thermometers, 793. 

thermo-electric pyrometer, 794. 
Brown coal, 23. 
Bucket conveyors, 225. 
Bucket traps, 659. 
Buckeye skimmer, 146. 






INDEX 



973 



Builder's rating of boilers, 120. 
Bundy steam separator, 648. 
Bundy steam trap, 660. 
Bunkers, coal, 223. 
Burke's smokeless furnace, 164. 
Burners, oil, 70-75. 

powdered coal, 60-66. 
Burning points, oils, 756. 
Bursting strength of pipes, 680. 

Cable-car coal conveyors, 236. 

Caking coals, 20. 

Calorific value of coals, 16-25, 32. 

Calorimeters, fuel, 808. 

Calorimeters, steam, 806. 

Capacity, effect of, on boiler efficiency, 

130. 
Carbon dioxide, percentage of, in flue 
gases, 42. 

properties of, 33. 
Carbon-hydrogen ratio, 16. 
Carbon monoxide, heat losses due to 
formation of, 49. 

properties of, 33. 
Carbon packing, Curtis turbine, 420. 
Carnot cycle, 308. 

Carpenter separating calorimeter, 807. 
Carrier, psychrometric chart, 527. 
Cast-iron pipes, 682, 684. 
Central condensing systems, 508. 
Central hydrostatic cylinder lubricator, 

766. 
Central Station statistics, 852. 
Centrifugal oilers, 760. 
Centrifugal pumps, 629. 

characteristics of, 633. 

multi-stage, 632. 

performance of, 634. 

tables of sizes, 636. 

types of, 629. 

volute, 630. 
Centrifugal steam separators, 647. 
Chain grates, 172. 
Chattanooga Electric Company, test of 

spray fountains, 519. 
Check valves, 740. 

Chemical purification of feed water, 543. 
Chicago settings, hand-fired furnace, 159. 
Chimney at Armour Institute, 268. 
Chimney draft, 247. 



Chimney draft, table of, for various 

temperatures, 250. 
Chimney, efficiency of, 282. 

height of, for burning fuel oil, 259. 

material for brick, 271. 

self-sustaining steel, 261. 

stability of brick, 272. 

stability of steel, 264. 

steel, 261. 

strain sheet, for reenforced concrete, 
275. 

test of an unlined steel, 254. 

test of 100-foot steel, 254. 

thickness of shell, steel, 261. 

thickness of walls, brick, 267. 

Weber, reenforced concrete, 275. 
Chimney gases, analysis of, 47. 
Chimney vs. mechanical draft, 300. 
Chimneys, 247-284. 

brick, 265. 

classification of, 259. 

concrete, 275. 

core and lining for, 271. 

cost of, 283. 

Custodis radial brick, 266, 275. 

dimensions of, 258, 284. 

formulas for, 252. 

foundations for, 281. 

guyed steel, 260. 
Cincinnati Traction Company, coal con- 
veyors, 235. 
Circulating pumps, 640. 
Classification of boilers, 89. 

chimneys, 259. 

condensers, 466. 

feed-water heaters, 554. 

fuel-oil burners, 70-75. 

fuels, 15. 

powdered-coal burners, 59. 

pumps, 588. 

steam separators, .646. 

steam traps, 657. 

steam turbines, 386. 

stokers, 171. 

testing instruments, 772. 
Clausius cycle, 310. 
Cleaners, tube, 150. 

Clearance volume, influence of, on en- 
gine economy, 331. 
Coal, 9-16. 



974 



INDEX 



Coal, analysis of, 15-25. 

anthracite, 18. 

bituminous, 20-24. 

calorific value of, 32. 

classification of, 16. 

composition of, 16. 

heating value of, 32. 

powdered, 58-66. 

proximate analysis, 32. 

purchasing, 56, 880. 

sampling, 917. 

size of bituminous, 54. 

specifications for purchasing, 880. 

storage of, 222. 

ultimate analysis, 32. 

washed, 54. 
Coal and ash handling, 222. 
Coal bunkers, 223. 
Coal conveyors, 224. 
Coal fields of the United States (ref.), 22. 
Coal hoppers, 242. 
Coal storage, 131. 
Coal valves, 244. 
Cochrane heater, 555. 
Coefficient of expansion, pipe materials, 

694. 
Cold tests, oil, 756. 
Columbia expansion trap, 661. 
Combustion, 29, 122. 

ilameless, 30. 

products of, 38. 

rate of, 122. 

surface, 30. 

temperature of, 21. 
Commercial National Bank Building, 

Chicago, Ash System, 232. 
Commonwealth Edison Company, Fisk 

Street Station, 885. 
Compound engines, 350. 
Compounds, boiler, 543. 
Compressed-air oiling system, 762. 
Compressed air, power required, air lift, 

642. 
Compression, effect of, on engine econ- 
omy, 332. 
Concrete chimneys, 275. 
Condensation and leakage losses in en- 
gines, 327. 
'Condensers, 461-537. 

air for cooling purposes, surface, 500. 



Condensers, Alberger barometric, 479. 
Baragwanath siphon, 475. 
Baragwanath surface, 482. 
barometric, 477. 
Blake jet, 469. 
choice of, 516. 
classification of, 466. 
cooling water for jet, 470. 
cooling water for surface, 491. 
cost of, 262. 
counter-current, 478. 
Dalton's Laws, 461. 
dry-air surface, 499. 
dry tube, 485. 

economical vacuum for, 515. 
ejector, 476. 

extent of cooling surface for, 492. 
function of, 464. 
high vacuum, 484. 
independent, 505. 
injection orifice, 474. 
jet, 469. 

Korting multi-jet, 491. 
location of, 505. 
low head, 481. 
multi-flow surface, 484. 
Pennel saturated air, 502. 
power consumption of auxiliaries, 510. 
Schutte ejector, 476. 
siphon, 474. 
sizes of siphon, 475. 
specifications for, 869. 
standard low vacuum, 467. 
surface, dry-air cooled, 499. 
surface, evaporative, 501, 503. 
surface, water-cooled, 475. 
tests of evaporative, 504. 
tests of surface, 492. 
Tomlinson barometric, 480. 
volume of condenser chamber for jet, 

474. 
Weighton multi-flow, 494. 
Weiss barometric, 478. 
Westinghouse-Leblanc, 489. 
Wheeler admiralty, 483. 
Wheeler low-head centrifugal jet, 481. 
Worthington barometric, 480. 
Worthington jet, 468. 
Condensing, influence on engine econ- 1 

omy, 361. 









INDEX 



975 



Condensing plant, elementary, 2. 
Condensing plant with full complement 

of heat-saving appliances, 13.. 
Convection, heat transmitted by, 115. 
Conversion tables, 963. 
Conveyors, 224. 
Cooling ponds, 517. 
Cooling towers, 520. 
Cooling towers, fan vs. natural draft, 

523. 
Copper pipes, 678. 
Corliss engine, 315, 350. 
Correction factors for steam turbines, 

450. 
Cost, boilers and settings, 141. 

chimneys, 283. 

condensers, 512. 

engines, 385. 

evaporating water, 133. 

handling coal and ashes, 242. 

initial, of steam engine plants, 822. 

initial, of steam turbine plants, 823. 

mechanical draft systems, 289. 

pipe flanges, 686. 

power (see Power costs). 

pulverizing coal, 65. 

stokers, 189. 

turbines, 453. 
Costs, operating, 830. 
Coverings for steam pipes, 692. 
Crusher and cross conveyor, 228. 
Curtis steam turbine, 412. 
Curve load factor, 815. 
Custodis radial brick chimney, 266. 
Cut-off, commercial, 931. 
Cut-off, point of, 931. 
Cylinder condensation, 327. 
Cylinder cups, 764. 
Cylinder efficiency, 325. 
Cylinder lubrication, 764. 
Cylinder ratios, compound engines, 350. 

Daily load curves, 856. 
Daily power-house report, 818. 
Dalton's laws, 461. 
Damper regulators, 146. 
Davis back-pressure valve, 746. 
Dean air pump, 618. 
De Laval centrifugal pump, 637. 
steam turbine, 390. 



Delray Station, large boiler at, 112. 
Demand factors, 816. 
Density of air and flue gas, 249. 
Depreciation of powdered-coal furnace, 

65. 
Depreciation, rate of, 823. 
Desmond injector, performance of, 614. 
Detroit Edison Company, coal-handling 

system, 236. 
Diagram factor, steam engine, 318. 
Diaphragm valve, 726. 
Differential traps, 663. 
"Direct" steam separator, 647. 
Disk-valve pump, 593. 
Disk-water meter, 778. 
Divergent nozzle, design of, 395. 
Double-flow steam turbine, 433. 
Double stoker, 137, 179. 
Down-draft furnace, 164. 
Draft and rate of combustion, 251. 
Draft, balanced, 306. 

chimney, 247. 

fan, 289. 

for powdered coal burners, 65. 

forced, 289. 

gauges, 791. 

induced, 289. 

influence of, on boiler efficiency, 117. 

influence of, on locomotive boiler, 288. 

influence of, on torpedo boat boiler, 
117. 

mechanical, 285-289. 

through boiler settings, 250. 
Drainage of jackets and receivers, 665. 
Drains, office building, 672. 
Drips, 655. 

high-pressure, 657. 

low-pressure, 655. 

removal of oil from, 671. 

under alternate pressure and vacuum, 
668. 

under vacuum, 667. 
Dry-air pumps, 622. 
Dry-air surface condensers, 499. 
Dry docks, centrifugal pump character- 
istics for, 633. 
Dry tube surface condensers, 485. 
Dulong's formula, 34. 
Dunham steam trap, 662. 
Dunham vacuo-vapor system, 674. 



976 



INDEX 



Duplex coal valve, 245. 
Duplex steam pump, 590. 
Duplicate piping system, 703. 
Dutch ovens, 156. 
Duty, pump, 603. 
Dynamometers, 800. 

Economizers, 576. 

factors for determining installation of, 
578. 

green, 576. 

heat transmission in, 578. 

tests of, 581. 

value of, 577. 
Edwards air pump, 622. 
Efficiencies of boilers and grates, 123. 

boilers with oil fuel, 71. 

boiler-feed pumps, 597, 602. 

centrifugal pumps, 634. 

compound engines, saturated steam, 
360. 

compound engines, superheated steam, 
372. 

fans, 293. 

piston pumps, 597, 602. 

simple engines, saturated steam, 348. 

simple engines, superheated steam, 
372. 

steam turbines, 448. 

triple-expansion engines, superheated 
steam, 375. 

triplex pumps, 611. 
Efficiency and capacity, influence of oil 

fuel on boiler, 137. 
Efficiency, air lift, 642. 

Carnot cycle, 308. 

commercial, 1. 

condensing plants, 7. 

cylinder, 325. 

fuel oil furnaces, 69. 

furnace, 123. 

grate, 124. 

mechanical, of steam engines, 323. 

non-condensing plants, 2. 

powdered coal furnaces, 64. 

Rankine cycle, 310. 

ration of engine, 323. 

standards for steam engines, 319. 

thermal, of steam engines, 322. 
Ejector condenser, 476. 



Ejector, Shone, 672. 

Electrical power, cost of (see Power 

costs). 
Elementary condensing plant, 7. 
Elementary non-condensing plant, 2. 
Elementary theory, Curtis turbines, 
423. 

De Laval turbine, 394. 

Kerr turbine, 411. 

Terry turbine, 407. 

Westinghouse-Parsons turbine, 436. 
Elevating tower, cable-car distribution, 

236. 
Elevating tower, hand-car distribution, 

233. 
Ellison's universal calorimeter, 807. 
Emergency valves, 739. 
Engines (steam), 307-385. 

A.S.M.E. code for testing, 927. 

automatic cut-off, 364. 

back-pressure, effect of, on economy, 
334. 

binary vapor, 379. 

Carnot cycle, 308. 

Clausius cycle, 310. 

clearance volume, 331. 

compound, 350. 

compression, effect of, on economy, 
361, 363. 

Corliss, 315. 

cost of, 385. 

cylinder condensation, 327. 

economy of (see Tests). 

efficiencies of (see Efficiencies). 

friction of, 355. 

heat consumption of, 320. 

high-speed, 342, 347. 

ideal, 307. 

incomplete cycle, 313. 

incomplete expansion, loss due to, 
332. 

increasing initial pressure, effect of, 
337. 

increasing rotative speed, 341. 

jackets, influence of, 340. 

leakage losses in, 329. 

Lentz superheated steam, 371. 

locomobile, 376. / 

logarithmic diagram for, 329. 

low-speed, 349. 









INDEX 



977 



Engines (steam), methods for increasing 
economy of, 337. 

mechanical efficiency, 323. 

moisture in steam, 336. 

multi-valve, 347. 

pumping, economy of, 359. 

Rankine cycle, 310. 

receiver-reheaters, economy of, 339. 

rotary, 382. 

simple, 342. 

single-valve, 342. 

specifications, 861. 

steam consumption of, 320. 

straight-flow, 378. 

Sulzer, 370. 

superheated steam, 367. 

tests of (see Tests). 

thermal efficiency, 322. 

throttling vs. automatic cut-off, 364. 

triple and quadruple expansion, 359. 

uniflow, 378. 

wire-drawing, effect of, 334. 

with low-pressure turbines, 441. 
Entropy diagram, 308-312. 
Equation of pipes, 719. 
Exhaust heads, 654. 
Exhaust piping, 721. 
Expansion of pipe materials, 694. 
Expansion, ratio of, 318. 
Expansion traps, 660. 
Extended boiler front setting, 97. 
Evaporation, cooling pond, 517. 

cost of, oil vs. coal, 70. 

from and at 212° F. per square foot 
per hour, 120. 

rate of, boilers, 127. 

rates of, in still air, 517. 

unit of, 114. 
Evaporative surface condenser, 501, 
503. 

Factor of evaporation, 114. 

Fan draft, 289. 

Fans, capacity of forced draft, 301. 

capacity of induced draft, 300. 

capacity of Sirocco blowers, 302. 

performance of, 292, 

theory of, 292. 
Feed-water, analysis of, 540. 
Feed-water heaters (see Heaters). 



Feed-water heating system, choice of, 

582. 
Feed-water piping, 727. 
Feed-water purification, 543-547. 
Feed- water regulators, 607. 
Feed- water temperature, ideal, 321. 
Feed-water, weighing of, 774. 
Fery radiation pyrometer, 795. 
Filters, oil, 769. 
Fire, thickness of, 135. 
Fire-box boilers, 91. 
Fire-tile combustion chamber, 173. 
Fire-tube boiler, 91. 
First National Bank Building, Chicago, 

power costs, 819. 
Fisher pump governor, 607. 
Fittings, pipe, 678. 
Fitzgibbons boiler, 92. 
Fixed carbon in coal, 16. 
Fixed charges, 820. 
Flameless combustion, 86. 
Flanged fittings, 683. 
Flanges, table of standard, 686. 
Flap coal valve, 245. 
Flash point, oil testing, 756. 
Flemming four-valve engine, 358. 
Flinn trap, 663. 
Float trap, 658. 
Floor space, turbine vs. reciprocating 

engine, 446. 
Flow of steam in pipes, 710. 

steam through nozzles, 397. 

water through pipes, 730. 
Flue gas analysis, 47, 801. 
Flue gas apparatus, 802-806. 

Little, modified, 803. 

Orsat, 802. 

Simmance-Abady recorder, 804. 

Uehling composimeter, 805. 

Williams improved, 802. 
Flue gas, effect of forcing boiler on tem- 
perature of, 134. 
Flue gas, heat loss in dry, 45. 

temperatures, 134. 
Flush-front boiler setting, 98. 
Flywheel pumps, 598. 
Foot valves, 749. 
Forced draft, 290. 
Forced-feed lubricator, 766. 
Forcing capacity of boilers, 134. 



978 



INDEX 



Foster back-pressure valve, 746. 

pressure regulator, 748. 
Foster superheater, 201. 
Foundations, chimney, 281. 
Fountain, spray, 518. 
Four-valve engines, tests of, 357. 
Friction of engines, 335. 
Friction of water in pipes, 719, 732. 
Friction tests of oil, 756. 
Friction through valves and fittings, 719. 
Fuel calorimeters, 808. 
Fuel, cost of, 833. 
Fuel oil, 66-80. 
Fuel oil burners (see Burners). 
Fuel ratio, 16. 
Fuel weighing, 772. 
Fuels and combustion, 16-29. 
Fuels, classification of, 16. 
Function of the condenser, 464. 
Furnace arch bars, 101. 
Furnace efficiency, 123. 
Furnace for burning oil fuel, 76, 77. 
Furnace for burning powdered coal, 59. 
Furnace, hand-fired, 154. 
Furnace influence on gas composition, 49. 
Furnace temperature, influence on boiler 

efficiency, 138. 
Furnaces (see Smokeless furnaces). 
Fusible plugs, 149. 

Gaseous fuels, 85. 
. characteristics of, 87. 

Gate valves, 736. 

Gauge cocks, 148. 

Gauges, water, 148. 

Gebhardt steam meters, 783. 

Geipel steam trap, 661. 

General Electric Company steam meter, 
785. 

Globe valves, 736. 

Goubert feed-water heater, 560. 

Goutal's formula, 35. 

Government specifications for purchas- 
ing coal, 880. 

Governor steam pump, 607. 

Grate bars, thickness of, 143. 

Grate, loss of fuel through, 49. 

Grate surface, 121. 

Grates, chain, 178. 
shaking, 144. 



Grates, stationary, 142. 
Gravity oil feed, 761. 
Gravity, Baume, oils, 755. 
Gravity, specific, oils, 755. 
Grease extractor, 653. 
Greases, 752. 
Green chain grate, 172. 

economizer, 576. 
Gruner's classification, 17. 
Gutermuth valves, 623. 
Guyed steel chimney, 260. 

Hamler-Eddy smoke recorder, 188. 

Hammel fuel oil burner, 73. 

Hammel oil-burning furnace, 76. 

Hammer tube cleaner, 151. 

Hancock injector, 613. 

Hand-fired furnace, 154. 

Hand shoveling, 224. 

Hangers, for pipes, 697. 

Hardness of feed- waters, 543. 

Hartford Boiler Insurance Company, 

annual report, 541. 
Hartford boiler specifications, 865. 
Hawley down-draft furnace, 164. 
Headers, main steam, 707. 
Heat balance, boiler tests, 52. 
Heat consumption of engines, 320. 
Heat distribution, condensing plants, 10, 

13. 
Heat distribution, non-condensing plants, 

4,7. 
Heat losses, in burning coal, 45-57. 

in the chimney gases, 45. 

in steam engines, 327. 
Heat of combustion, 31. 
Heat transmission, boilers, 115. 

closed heaters, 563. 

condensers, 492. 

economizers, 578. 

influence of scale on, 539. 

superheaters, 212. 
Heater and purifier combined, 557. 
Heaters, feed-water, 538-587 

Baragwanath, 563. 

choice of, 570, 582. 

classification of, 554. 

closed, 560. 

Cochrane, 555. 

counter-current, 561. 






INDEX 



979 



Heaters, flue gas, 576. 

Goubert, 560. 

Harrisburg, 561. 

Hoppe's, 557. 

induced, 572. 

live steam, 572. 

open, 555. 

Otis, 562. 

parallel-current, 560. 

primary, 554. 

secondary, 554. 

single-flow, 560. 

steam tube, 562. 

temperatures in open, 557. 

through, 571. 

vacuum, 573. 

Wainwright, 551. 

Webster, 556. 
Heating surface, boilers, 118. 
Heine boiler, 106. 
Heine superheater, 204. 
Heinrich smokeless furnace, 169. 
Heintz expansion trap, 662. 
Herrick rotary engine, 382. 
Herringbone grate, 66. 
Hewes and Phillips air pump, 620. 
High- and low-speed engines, 342-349. 
High-pressure drips, 657. 
High-speed multi-valve engines, 347. 
High-speed single-valve engines, 342. 
High-vacuum systems, 486. 
Hoist and trolley coal-handling system, 

236. 
Hollow bridge wall, 168. 
HoUy loop, 669. 
Hoope's feed-water heater, 557. 

steam separators, 646. 
Horizontal return tubular boilers, 96. 
Horse power of boilers, 120. 
Hot-well pumps, 641. 
Hot-well temperatures, surface con- 
densers, 492. 
Hunt coal conveyor, 230. 
Hydraulic Oil Storage Company's fuel 

oil system, 83. 
Hydraulic packing, 596. 
Hydraulic valve gear, Curtis turbine, 

419. 
Hydrogen, losses due to, 51. 

properties of, 31, 33. 



Hydrometer, Baume, oils, 755. 
Hydrostatic cylinder lubricator, 765. 
Hygrometry, 523. 

Ideal engine, 307. 
Ideal feed- water temperature, 321. 
Identification of oils, 755. 
Illinois Engineering Company's Auto- 
matic vacuum valve, 723. 
Impulse turbine, 387. 
Incomplete combustion, loss due to, 47. 
Incomplete expansion, loss due to, 332. 
Increasing boiler pressure, economy of, 

337. 
Increasing degree of vacuum, economy 

of, 362. 
Increasing rotative speed, economy of, 

341. 
Independent condensers, 505. 
Independently fired superheaters, 205. 
Indicated horse power, 318. 
Indicator cards, air pump, 629.] 
Indicator cards, analysis of, 931. 
automatic cut-off engine, 366. 
Indicator cards, throttling engine, 364. 
Westinghouse-Parsons turbine, 431. 
Induced draft, 291. 
Induced heaters, 572. 
Inherent losses in burning fuel, 53. 
Initial condensation, 327. 
Initial temperature, influence of, 138. 
Injection orifice, 473. 
Injectors, 612. 
automatic, 613. 
performance of, 614. 
positive, 612. 

range in working pressures, 616. 
vs. steam pumps as boiler feeders, 617. 
Interest charges, 822. 
Intermittent oiling, 757. 
International Gas Company's fuel oil 

system, 82. 
Isolated stations, cost of power in (see 

Power, cost of). 
Isolated stations, influence of load factor 

on economy, 853. 

Jackets, influence of, 340. 
Jackets, methods of draining, 665. 
Jet, Bloomsburg, 286. 



980 



INDEX 



Jet, ring steam, 286. 

Jet condensers, 469. 

Jets, steam, 285. 

Jets, steam consumption of, 287. 

Jones underfeed stoker, 182. 

Kennicott feed-water purifier, 776. 
Kennicott water weigher, 775. 
Kent's fuel classification, 17. 
Kent's wing-wall furnace, 163. 
Kerosene, use of, in boilers, 544. 
Kerr steam turbine, 407. 
Keystone separator, 647. 
Kieley reducing valve, 748. 
Kindling temperatures, 29. 
Kirk wood oil burners, 74. 
Kitts feed- water regulator, 608. 
Kitts hydraulic damper regulator, 147. 
Knowles triplex pump, test of, 610. 
Korting fuel oil burner, 72. 
Korting multi-jet condenser, 491. 

Labor, cost of, in power plants, 830. 

cost of, in street railway plants, 832. 

cost of, in tall office buildings, 833-850. 
Labyrinth packing, 433. 
Leakage of steam in engines, 329. 
Leblanc air pump, 624. 
Lentz superheated steam engine, 371. 
Leyland automatic cylinder cup, 764. 
Life of power-plant appliances, 825. 
Lignite, 23. 

Limit of superheat, 192. 
Link Belt Company coal-handling sys- 
tem, 226. 
Little flue gas apparatus, 803. 
Live-steam feed-water heaters, 575. 
Live-steam separators, 644. 
Load curves, typical, 855-857. 
Load factor, 815. 

influence of, on cost of power, 815. 
Location of condensers, 505. 

of separators, 649. 

of traps, 664. 
Locomobile, 376. 
Loew grease extractor, 653. 
Logarithmic diagram for engines, 329. 
Loop header, 704. 
Loop, HoUy, 669. 
Loop, steam, 668. 



Loss of heat from bare pipes, 691. 

Loss of heat from covered pipes, 692. 

Losses in burning fuel, 45. 

Losses in steam engines, 327. 

Lower heating values of gasolene fuels, 

31. 
Low-pressure drips, 655. 
Low-pressure turbines, 439. 
Low-speed engines, 349. 
Lubricants, 750. 
Lubricating oils, classification of, 750. 

properties of, 754. 

specific gravity of, 755. 
Lubrication, atmospheric, 757. 

central hydrostatic system, 766. 

compressed-air feed, 762. 

cylinder lubrication, 764. 

forced system, 766. 

gravity systems, 761. 
Lubricators, hydrostatic, 765. 
Luckenbach smokeless furnace, 170. 
Luckenbach superheater, 208. 
Ludlow angle valve, 738. 
Lunkenheimer sight-feed lubricator, 765. 
Lyons boiler, 102. 

Mahler bomb calorimeter, 808. 

Mains, steam, 707. 

Maintenance, 829. 

Manning vertical boilers, 91. 

Manufacturer's 1912 schedule of flanges 
and fittings, 686. 

Mark's and Davis' steam tables, 961. 

Marsh boiler feed pump, 594. 

Marsh gas, properties of, 33. 

Marsh steam pump, test of, 600. 

Materials for brick chimneys, 271. 

Materials for pipes and fittings, 678 

Materials for superheaters, 209. 

McClaves argand blower, 286. 

McDaniel float trap, 658. 

Mean specific heat of gases, 44. 

Mean specific heat of superheated 
steam, 196. 

Mean temperature difference, heaters, 
578. 

Measurement of feed-water consump- 
tion, engine, 320. 

Measurement of heat units consumed by 
engines, 320. 



INDEX 



981 



Measurement of steam used by auxil- 
iaries, 510. 

Mechanical boiler-tube cleaner, 150. 

Mechanical draft, 285-306. 

Mechanical efficiency of engines, 323. 
of pumps, 597. 

Mechanical purification of feed water, 
546. 

Mechanical stokers, 170. 

Mechanical valve gear, Curtis turbine, 
418. 

Medium- and low-speed engines, 349. 

Mesh steam separators, 648. 

Meters, steam, 781. 

Meters, water, 777. 

Meyers bagasse furnace, 27. 

Meyers sawdust furnace, 28. 

Mineral oils, 751. 

Mixed pressure turbines, 439. 

Moisture in air, loss due to, in combus- 
tion, 50. 
in fuel, loss due to, 50. 
in steam, determination, 806. 
in steam, effect on engine economy, 

336. 
in steam evaporated by throttling, 365. 

Mollier diagram, 966. 

Mullan valveless air pump, 622. 

Murgue's theory, centrifugal fans, 292. 

Murphy furnace, 181. 

Napier's rule for the flow of steam, 

397. 
Nash water meter, 778. 
Natural-draft cooling tower, 520. 
Natural gas, properties of, 85. 
Nitrogen, properties of, 33. 
Non-condensing engines, test of, 347- 

350. 
Non-condensing plants, arrangement of, 
2,6. 
exhaust piping in, Paul system, 725. 
exhaust piping in, Webster system, 

721. 
feed-water piping in, 728. 
open heater in, 573. 
Non-return valves, 738. 
Nordberg uniflow engine, 378. 
Norfolk Traction Co.'s ash-handling 
system, 241. 



Northwestern Elevated R.R. power 

house, condenser piping, 511. 
Nozzles, De Laval steam turbine, 392. 
Nozzles, flow of steam through, 395. 

flow of water through, 730. 

Kerr steam turbine, 409. 

theoretical design of divergent, 395. 
Nugent telescopic oiler, 759. 

Office buildings, cost of power in, 843. 

Oil, atomization of, 79. 

Oil bath, 757. 

Oil burners, 72-75. 

branch, 74. 

Hammel, 73. 

Kirkwood, 74. 

Korting, 72. 

Warren, 75. 

Williams, 74. 

tests of, 78. 
Oil cups, 759. 

eliminators, 650. 

feeding system, 79. 

filters, 769. 

fuels, analysis of, 18-26. 

pressure in fuel-oil systems, 79. 

purchase of fuel, 83. 

separators, 650. 

storage, 82. 

transportation, 82. 
Oil, waste, and supplies, cost of, in 

power plants, 834. 
Oiler, centrifugal, 760. 

gravity, 761. 

pendulum, 761. 

ring, 760. 

telescope, 759. 
Oiling systems (see Lubricating sys- 
tems). 
Oils, animal, 750. 

identification of, 755. 

mineral, 751. 

properties of, 758. 

purchase of fuel, 83. 

specifications for, 758. 

specific gravity of, 755. 

tests for, 756. 

vegetable, 750. 
Olefiant gas, properties of, 33. 
Open heater, 555. 



982 



INDEX 



Open heater vs. closed heater, 570. 

Operating costs, 830. 

Operating costs, reciprocating engine vs. 

steam turbine, 837. 
Optical pyrometers, 795. 
Orifice measurements, 780. 
Orifice, size of injection, 473. 
Orifices, flow of steam through, 397. 

flow of water through, 730. 
Orsat apparatus, 802. 
Otis feed-water heater, 563. 
Output and load factor, 815. 
Overhead storage, bucket hoist, 235. 
Overload capacity, steam turbines, 447. 
Oxygen, properties of, 33. 

Pan surface, open heaters, 559. 

Parallel-current condenser, 468. 

Parallel-current feed-water heater, 560. 

Parker boiler, 107. 

Parr coal calorimeter, 809. 

Parson smokeless furnace, 168. 

Parsons vacuum augmenter, 488. 

Paul exhauster, 725. 

Paul heating system, 725. 

Peabody oil burner, 73. 

Peabody oil furnace, 77. 

Peat, 24. 

Pemberthy injector, 613. 

Pendulum oiler, 761. 

Pennel saturated-air surface condenser, 

502. 
Permanent statistics, power plant, 811. 
Pinther powdered-coal burner, 59. 
Pipe anchors, 697. 
Pipe bends, 695. 
Pipe, brass, 678. 

cast iron, 678. 

cast steel, 678. 

copper, 678. 

coverings, 692. 

mild steel, 677. 

sizes of standard, 681. 

threads, 680. 

wrought iron, 677. 
Pipe-flanges, 686. 

size of extra heavy, 686, 688. 

sizes of standard, 686, 688. 
Pipe hangers, 697. 
Pipe supports, 697. 



Pipe threads, United States standard, 

680. 
Pipes, equation of, 719. 

expansion of, 694. 

flow of steam in, 710. 

flow of water in, 730. 

friction in, 719. 

loss of heat from bare, 691. 

materials for, 676. 

size of, for low-pressure drips, 655. 

strength of, 679. 
Piping, auxiliary header system, 703. 

Commonwealth Edison Company, 
Fisk Street Station, 885. 

Des Moines City Railway Company, 
702. 

duplicate system of, 703. 

feed-water, 727. 

high-pressure, 698. 

La Salle Hotel, Chicago, 708. 

loop header system of, 704. 

Manhattan Elevated, New York, 
707. 

Paul heating system, 725. 

Princeton University, 700. 

specifications for, 871. 

spider system, 701. 

steam, 470-483. 

Webster heating system, 721. 

Yonkers power house, 705. 
Pistons, water, 595. 
Pitot tubes, 292, 786. 
Plungers, pump, 595. 
Ponds, cooling, 517. 
Pop safety valves, 744. 
Positive injectors, 612. 
Powdered coal, 58-66. 
Powdered coal burners, 61-63. 

Blake, 62. 

Pinther, 59. 

Schwartzkopff, 61. 

triumph, 63. 
Power, bibliography, 858. 

Boston Elevated, 846. 

central stations, 845-860. 

compound engine plants, 835. 

cost of, 810-860. 

demand factors, 816. 

depreciation, 823. 

Fall River, 849. 



INDEX 



983 



Power, First National Bank Building, 
Chicago, 819. 

Fitchburg Yarn Company, 851. 

fixed charges, 820. 

fuel, 833. 

Hyde Park Electric Light Co., Bos- 
ton, 851. 

insurance, 829. 

interest, 822. 

isolated stations, 842-845. 

labor, 830. 

large central stations, 848-890. 

maintenance, 829. 

obsolescence, 824. 

office buildings, 843. 

oil, waste and supplies, 834. 

operating charges, 830. 

operating records, 812. 

Pacific Gas & Electric Co., 849. 

permanent statistics, 812. 

simple engine plant, 835. 

steam turbine plants, 837. 

street railway plants, C. C. Moore, 
838-841. 

street railway plants, R. C. Carpenter, 
846. 

street railway plants, typical U. S., 
845. 

taxes, 829. 
Power consumption of condenser auxil- 
iaries, 510. 
Power measurements, 798. 
Power-plant design, elements of, 853. 
Power plants, A.S.M.E. code for test- 
ing, 943. 
Power plants, distribution of heat losses 
in, 1-15. 

elementary, 1-15. 
Powers thermostat, 726. 
Preheating feed water, economy of, 552. 
Pressure gauges, 792. 
Pressure of aqueous vapor for different 

temperatures, 465. 
Pressure regulator, 748. 
Preventable losses in burning fuel, 53. 
Products of combustion, 38. 
Properties of air, 524. 

of fuel oil, 67. 

of gases, 87. 

of lubricating oils, 758. 



Properties of steam, 961. 

Proximate analysis, 33. 

Psychrometric chart, 527. 

Pulsometer, 641. 

Pump governors, 607. 

Pumping engines, economy of, 359. 

Pumps, air, 618. 

air, jet condensers, 619. 

air, surface condensers, 626. 

air, theoretical work, 628. 

air lift, 642. 

boiler-feed, 590. 

centrifugal, 629. 

classification of, 588. 

circulating, 640. 

direct-pressure, 641. 

double suction centrifugal, 631. 

Duplex, 590. 

duty of, 603. 

effect of piston speed on economy, 599. 

field for, 631. 

fly-wheel, 598. 

hot-well, 641. 

jet, 612. 

Marsh boiler-feed, 594. 

multi-stage centrifugal, 632. 

outside packed plunger, 598. 

performance of piston, 597, 602. 

plungers for, 595. 

power, 609. 

pulsometer, 641. 

rotary, 638. 

simplex, 594. 

size of boiler-feed, 605. 

tests of (see Tests). 

triplex, 609. 

turbine, 630. 

vacuum, 618-627. 

volute, 630. 

water pistons for, 595. 
Purchasing coal, 56, 880. 

government specifications for, 880. 
Purification, chemical, feed- water, 543. 

mechanical, 546. 

thermal, 547. 
Purifiers, live steam, 575. 
Purifying plants, Anderson, 551. 

Kennicott, 548. 

Scaife, 550. 

We-Fu-Go, 551. 



984 



INDEX 






Pyrometers, air recording, 793. 
Bristol thermo-electric, 793. 
Callender resistance, 794. 
Fery radiation, 795. 
Wanner optical, 795. 

Qualification for a good lubricant, 754. 

Radial brick chimneys, 266. 

Radiation and minor losses, boilers, 52. 

Radiation, heat transmitted by, 115. 

Radiation pyrometer, 795. 

Radiator drains, 672. 

Rankine cycle, 310. 

Rate of combustion, powdered coal, 64. 

Rate of combustion, relation of air sup- 
ply to, 304. 

Rate of depreciation, 823. 

Rate of driving, effect on economy of 
boilers, 130-138. 

Rateau regenerator accumulator, 443. 
steam turbine installation, low pres- 
sure, 442. 

Rates of combustion, average, 122. 

Reaction turbines, 440. 

Receiver, reheaters, 339. 

Reciprocating engines vs. steam tur- 
bines, 441. 

Records, power plant, 810-815. 

Reducing valves, 747. 

Reenforced concrete chimneys, 275. 

Regenerator-accumulator, 443. 

Regulation of steam turbines, 447. 

Regulators, damper, 146. 
feed-water, 607. 

Repairs, cost of, power plants (see 
Power, cost of). 

Restricted feed, lubrication, 757. 

Return trap, 665. 

Returns tank, 671. 

Rhode Island power house ash-hand- 
ling system, 231. 

Ring oiler, 760. 

Ring steam jet, 286. 

Ringlemann smoke chart, 187. 

Riveted joints, steel chimneys, 264. 

Robb-Mumford boiler, 95. 

Robins belt conveyor, 232. 

Rochester forced-feed lubricator, 767. 

Roney stoker, 178. 



Rotary engines, 382. 
Rotary pumps, 638. 
Rowe feed-water regulator, 608. 

Safety valve, dead weight, 743. 

lever, 743. 

pop, 744. 
Safety valves, capacity of, 743. 
Sampling coal, 917. 
Saturated-air surface condenser, 501. 
Saturated-steam tables, 963. 
Sawdust as fuel, 25. 
Scaife system of feed-water purification, 

550. 
Scale, analyses of boiler, 540. 

influence on heat transmission, 539. 
"S-C" feed-water regulator, 609. 
Schmidt, independently fired superheat- 
ers, 206. 
Schutte ejector condenser, 477. 
Schwartzkopff powdered-fuel burner, 

61. 
Schwoerer superheater, 203. 
Scioto Valley Traction Co., coal-hand- 
ling system, 237. 
Scotch marine boiler, 93. 
Screwed fittings, 460. 
Seaton coal valve, 246. 
Sederholm boiler, 102. 
Seger cones, 796. 

Self-closing water gauge valves, 149. 
Separating calorimeters, 806. 
Separators, 646-653. 

Austin, 648. 

baffle plate, 648. 

Baum, 652. 

Bundy, 648. 

centrifugal, 647. 

classification of, 646. 

direct, 649. 

exhaust steam, 650. 

Hoppes, 646. 

Keystone, 647. 

live steam, 644. 

location of, 649. 

Loew, 653. 

mesh, 648. 

oil, tests of, 651. 

reverse current, 646. 

Stratton, 647. 



INDEX 



985 



Settings, smokeless, 154-182 

standard, 90-109. 
Shaking grates, 143. 
Shone ejector, 672. 

Simmance-Abady flue gas recorder, 804. 
Simplex coal valve, 244. 
Siphon condensers, 474. 
Siphon traps, 664. 
Skimmer, Buckeye, 146. 
Slip expansion joint, 697. 
Smoke determinations, 186. 
Smoke distribution in Chicago plants, 

154. 
Smoke prevention, 187. 
Smoke record chart, 187. 
Smoke recorders, 189. 
Smoke unit, 187. 
Smoke, visible, loss due to, 51. 
Smokeless furnaces, 153-173. 

balanced draft, 305. 

Burke's 164. 

Chicago settings, 159. 

down-draft, 164. 

Dutch ovens, 156. 

fire tile, 173. 

hand-fired, 154. 

Heinrich's, 168. 

Kent's wing-wall, 163. 

Luckenbach, 170. 

Parson, 168. 

steam jets for, 167. 

step grates, 178. 

stokers, 1 70-180. 

twin fire-arch, 157. 

Wooley, 163. 
Solid lubricants, 752. 
South Side Elevated Ry. Co., Chicago, 
chimney at, 262. 

coal crusher and cross conveyor at, 
228. 
Special furnaces, 153-170. 
Specific gravity of lubricating oils, 752. 
Specific heat of gases, 44. 
Specific heat of superheated steam, 195. 
Specific volume of superheated steam, 

198. 
Specifications, boiler, 865. 

condenser, 869. 

engine, 861. 

government, for purchasing coal, 880. 



Specifications, piping, 871. 
Speed, influence of piston on pump 
economy, 599. 

influence on engine economy, 341. 
Speed measurements, 799. 
Spider system of piping, 701. 
Spiro turbine, 460. 
Spray fountain, 518. 
Sprinkling stokers, 185. 
Stability of brick chimneys, 272. 
Stability of steel chimneys, 264. 
Standard methods of starting and stop- 
ping boiler tests, 599. 
Standby losses in boilers, 52. 
Station load factor, 815. 
Steam boilers, 89-152. 
Steam engines, 307-385. 
Steam, flow of, in pipes, 710. 
Steam jets, 167, 285, 287. 
Steam loop, 667. 
Steam mains, 707. 
Steam piping, 470-483. 
Steam, properties of, 963. 
Steam pumps, 588-605. 
Steam separators, 644-650. 
Steam, specific heat of superheated, 115. 
Steam traps, 657-667. 
Steam turbines, 386-480. 
Steel chimneys, 261. 
Steel-concrete chimneys, 275. 
Step bearing, Curtis turbine, 420. 
Stirling boiler, 110. 
Stirling superheater, 201. 
St. John's steam meter, 790. 
Stokers, 171-189. 

American underfeed, 183. 

Babcock & Wilcox, 173. 

chain grates, 172. 

cost of, 189. 

green chain grate, 176. 

Jones underfeed, 182. 

Murphy, 181. 

sprinkling, 185. 

Taylor, 183. 

underfeed, 183. 

Wilkinson, 180. 
Stop valves, 735. 
Storage, coal, 222. 

oil, 82. 

powdered coal, 65. 



986 



INDEX 



Straight flow engines, 378. 
Stratton separator, 647. 
Suction, maximum for pumps, 604. 
Sulphur dioxide, properties of, 33. 
Sulzer engine for superheated steam, 370. 
Superheat, economy of, 191. 
Superheat, limit of, 192. 
Superheated steam, 190-221. 

economy of, steam engine, 372. 

economy of, steam turbines, 448. 

properties of, 193. 

specific heat of, 195. 

specific volume of, 198. 
Superheaters, 194. 

Babcock & Wilcox, 200. 

Foster, 201. 

Heine, 204. 

independently fired, 205. 

Luckenbach superheater, 208. 

materials for, 209. 

Schmidt, 206. 

Schwoerer, 203. 

Stirling, 201. 

tests of independently fired vs. flue 
fired, 218. 

tests, miscellaneous, 216. 

yearly expense for repairs, cast-iron, 
211. 
Superheating moisture in air, loss due to, 

50. 
Superheating surface, extent of, 210. 
Supports, pipe, 697. 
Surface blow, 146. 
Surface combustion, 86. 
Surface condensers, 475. 
Surface, cooling, for condensers, 492. 
Surface, heating, for feed-water heaters, 

563. 
Surface, heating, for superheaters, 210. 
Suspended boiler setting, 99. 

Tachometers, 790. m 

Tan bark as fuel, 25. 

Tank, blow-off, 145. 

Tank, returns, 671. 

Taylor underfeed stoker, 183. 

Telescopic oiler, 759. 

Temperature drop throughout boiler 

setting, 140. 
Temperature-entropy diagram, 308-316. 



Temperature, influence of initial, on 

boiler efficiency, 138. 
Temperature measurements, 792. 
Temperature of combustion, 43. 
Temperature of flue gases, forced boiler 

capacity, 134. 
Temperature regulators, 726. 
Terry steam turbine, 405. 
Tesla bladeless turbine, 458. 
Tests of blowers, 296-298. 
Tests of boilers: 

A.S.M.E. code, 916. 

coal burning, 128. 

evaporation, Armour Glue Works, 133. 

evaporation, Delray boilers, 134. 

evaporation, Fisk Street Station, 132. 

influence of draft on efficiency, 130. 

influence of rate of combustion on air 
supply, 304. 

influence of rate of driving on capacity, 
135. 

influence of size of coal on capacity, 55. 

influence of thickness of fire on effi- 
ciency, 138. 

oil fuel, 71, 137. 

powdered coal, 64. 
Tests of burners, oil fuel, 78. 
Tests of chimney, 100-foot steel, 254. 
Tests of condenser auxiliaries, 510. 
Tests of condensers, 513. 

dry air, 501. 

evaporative, 504. 

Pennel saturated air, 503. 

Weighton surface, 492. 
Tests of cooling towers, 535. 
Tests of economizers, 581. 
Tests of engines, A.S.M.E. code, 927. 

Ball- Wood direct-connected set, 346. 

binary vapor, 379. 

compound condensing vs. non-con- 
densing, 354. 

compound vs. simple high-speed, 354. 

condensing, increase in power due to, 
363. 

Corliss, 354. 

5500- horse-power engine at Waterside 
Station, 355, 364. 

four-valve vs. single-valve high-speed, 
349. 

friction, 335. 






INDEX 



987 



Tests of engines, increasing back pres- 
sure, 334. 

increasing initial pressure, 338. 

locomobile, 376. 

Reeves simple engine, 346. 
Tests of engines, simple high-speed en- 
gines, saturated steam, 343. 

superheated steam, compound, 374. 

superheated steam, influence of degree 
of superheat, 369. 

superheated steam, influence of size of 
engine, 368. 

superheated steam, record perform- 
ance of, 377. 

superheated steam, triple, 375. 

tables of, 343, 360. 
Tests of furnace, relation of gas compo- 
sition to temperature, 48. 
Tests of injectors, 614. 
Tests of jets, steam, 287. 
Tests of oil burners, 78. 
Tests of oil separators, 651. 
Tests of oils, fuel, 71. 
Tests of oils, lubricating, 754. 
Tests of pipe coverings, 692. 
Tests of pumps, 610. 

air lift, 642. 

boiler feed, 600. 

centrifugal, De Laval, 637. 

centrifugal, Lea-Degan, 634. 

centrifugal, Worthington, 634. 

direct-connected triplex, 610. 

duplex fire pump, 599. 

geared triplex, 611. 

rotary, 639. 
Tests of separators, 645, 651. 
Tests of spray fountains, 519. 
Tests of superheaters, 218. 
Tests of turbines, 454. 
Thermal efficiency of engines, 322. 
Thermal purification of feed water, 547. 
Thermo-electric pyrometers, 793. 
Thermometers, classification of, 796. 
Thermostat, powers, 726. 
Thickness of fire, 135, 140. 
Thickness of shell, steel chimneys, 261. 
Thickness of walls, brick chimneys, 267. 
Throttling calorimeter, 806. 
Throttling, moisture evaporated by, 365. 
Throttling vs. automatic cut-off, 364. 



Tilden damper regulator, 148. 
Tile-roof furnaces, 173. 
Tomlinson condenser, 480. 
Towers, Alberger, 521. 

Barnard, 520. 

cooling, 520. 

test of, 536. 

theory of, 529. 

Worthington, 521. 
Traps (steam): 

acme, 659. 

bowl, 660. 

bucket, 659. 

Bundy, 660. 

classification of, 657. 

Columbia, 661. 

differential, 663. 

Dunham, 662. 

expansion, 660. 

Flinn,"663. 

float, 653. 

Geipel, 661. 

Heintz, 662. 

location of, 664. 

return of, 665. 

siphon, 664. 
Traveling coal hoppers, 243. 
Traveling grates, 172. 
Triumph powdered-coal furnace, 63. 
Try cocks, 3. 
Tube cleaners, 150. 
Tupper grate, 143. 
Turbine pumps, 630. 
Turbine tube cleaner, 151. 
Turbines (steam), 386-460. 

advantages of, 446. 

Allis-Chalmers, 435. 

A.S.M.E. code for testing,' 939. 

bleeder, 446. 

classification of, 386. 

correction factors, 450. 

cost of, 453. 

cost of operation, 455. 

Curtis, 412. 

De Laval, 390. 

double-flow, 433. 

economy of space, 446. 

efficiency of, 448. 

elementary theory, 388. 

impulse, 390. 



988 



INDEX 



Turbines (steam), influence of superheat, 
455. 

influence of vacuum, 456. 

Kerr, 407. 

low-pressure, 439. 

mixed pressure, 439. 

overload capacity, 447. 

Parsons, 428. 

reaction, 428. 

regulation, 447. 

simplicity of, 447. 

Spiro, 459. 

Terry, 405. 

Tesla bladeless, 458. 

tests of, 454. 

Westinghouse, 427. 

Westinghouse-Parsons, 428. 
Turner oil filter, 769. 
Twin fire-arch furnace, 157. 

Uehling composimeter, 805. 

Ultimate analysis, 33. 

Underfeed stokers, 182. 

Uniflow engines, 378. 

Unit of evaporation, 114. 

Units, conversion, 965. 

Universal calorimeter, 807. 

Useful life of power-plant appliances, 825. 

Vacua, increase of power due to increas- 
ing, 363. 
Vacua, influence of, on economy of en- 
gines, 363. 
Vacua, influence of high, on steam tur- 
bines, 456. 
Vacua-vapor system, 674. 
Vacuum ash conveyor, 238. 
Vacuum augmenter, Parsons, 488. 
Vacuum chambers, 594. 
Vacuum, corrections for standard con- 
ditions, 463. 
degree of, as affected by aqueous 

vapor, 473. 
drips under, 667. 
most economical, 515. 
Vacuum pumps, 618-625. 
Vacuum systems, high, 484. 
Valves, Acton atmospheric relief, 747. 
Anderson automatic non-return, 738. 
Anderson triple duty emergency, 739. 



Valves, atmospheric relief, 747. 

automatic non-return, 738. 

back-pressure, 746. 

blow-off, 741. 

check, 740. 

coal, 244. 

consolidated pop safety, 744. 

Crane atmospheric, 647. 

Crane hydraulic emergency, 739. 

Davis back-pressure, 746. 

diaphragm, 727. 

disk, 374. 

emergency, 739. 

foot, 749. 

Foster back-pressure, 746. 

gate and globe, 736. 

Gutermuth, 623. 

Illinois Eng. Company's vacuum, 723. 

Kieley reducing, 748. 

Ludlow angle, gate pattern, 738. 

non-return, 738. 

nozzle, Curtis turbine, 413. 

nozzle, De Laval turbine, 392. 

nozzle, theoretical, 395. 

Paul vacuum, 725. 

reducing, 747. 

safety, 743. 

stop, 735. 

Webster vacuum, 723. 
Van Stone joint, 684. 
Vegetable oils, 750. 
Velocity of steam through nozzles, 395. 

through pipes, 730. 
Velocity of water through nozzles, 730. 

through pipes, 730. 
Venturi meter, 778. 
Vertical blow-off connections, 145. 
Vertical tubular boilers, 90. 
Viscosity of oils, 756. 
Visible smoke, loss due to, 51. 
Volatile matter in coal, 15-25. 
Volume of jet condenser chamber, 474. 
Volute centrifugal pump, 630. 

Wainwright feed- water heater, 561. 

Wall brackets, piping, 697. 

Wanner optical pyrometer, 795. 

Warren fuel-oil burner, 75. 

Washed coals, 54. 

Water and boiler scale, analyses of, 540. 









INDEX 



989 



Water columns, 147. 

Water-cooling systems, 517. 

Water, flow of, in pipes, 730. 

Water, friction coefficient in clean iron 

pipes, 732. 
Water, height of lift by suction, 604. 
Water pistons, 595. 
Water-softening plants, 547. 
Water, weight of, for condensing, 470, 

491. 
Waterworks, centrifugal pump charac- 
teristics for, 633. 
Weber concrete-steel chimneys, 275. 
Webster feed-water heater, 556. 
Webster heating system, 721. 
Webster vacuum valve, 721. 
We-Fu-Go purifying system, 551. 
Weighing fuel, 772. 
Weighing water, 774. 
Weight of air as indicated by flue gas 

analysis, 42. 
Weight of boiler compound necessary, 

546. 
Weight of guyed steel chimneys, 260. 
Weight of water evaporated per square 

foot of heating surface, 120. 
Weighton multi-flow surface condenser, 

494. 
Weir measurements, 777. 
Weiss barometric condenser, 478. 
Westinghouse double-flow turbine, 433. 
Westinghouse impulse turbine, 427. 
Westinghouse-Leblanc condenser, 489. 



Westinghouse-Parsons steam turbines, 

429. 
Wet air pumps, 622. 
Wheeler admiralty surface condenser, 

483. 
Wheeler dry tube, 485. 
Wheeler low head centrifugal jet con- 
denser, 481. 
Wheeler, C. H., multi-flow surface con- 
denser, 484. 
rotrex vacuum pump, 623. 
White-star oil filter, 769. 
Wickes, boiler, 107. 
Wilcox water weigher, 776. 
Wilkinson stoker, 180. 
Williams flue gas apparatus, 802. 
Williams oil burner, 74. 
Wilsey relative efficiency meter, 797. 
Wing- wall furnace, 163. 
Wire drawing, 334. 
Wood as fuel, 25. 
Wooley smokeless furnace, 163. 
Worthington barometric condenser, 480. 
cooling tower, 521. 
jet condenser, 468. 
multi-stage centrifugal pump, 632. 
water meter, 777. 
weight determinator, 774. 
Wrought-iron pipe, 676. 

Yonkers power house, N.Y.C.R., piping 
for, 705. 

Zinc, use of, in boilers, 545. 






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